Control device of compression-ignition engine

ABSTRACT

A method of implementing control logic of a compression-ignition engine is provided. A control part of the engine performs a calculation according to the control logic corresponding to an engine operating state in response to a measurement of a measurement part, controls a fuel injection part, a variable valve operating mechanism, an ignition part and a supercharger so that a G/F becomes leaner than a stoichiometric air fuel ratio and a A/F becomes equal to or richer than the stoichiometric air fuel ratio, while causing the supercharger to boost, and controls the ignition part so that unburnt mixture gas combusts by self-ignition after the ignition. The method includes determining a supercharging pressure P, and determining control logic defining a close timing IVC of an intake valve. When determining the control logic, the close timing IVC (deg.aBDC) is determined so that the supercharging pressure P (kPa) satisfies the following expression: P≥8.0×10−11IVC6−1.0×10−8IVC5+3.0×10−7IVC4−4.0×10−6IVC3+0.0068IVC2−0.3209IVC+116.63.

TECHNICAL FIELD

The technology disclosed herein relates to a control device of acompression-ignition engine.

BACKGROUND OF THE DISCLOSURE

It is known that combustion by compressed self-ignition, in which amixture gas combusts instantly without flame propagation, maximizes fuelefficiency since the combustion period is minimal. However, variousproblems must be solved for automobile engines with regard to combustionby compressed self-ignition. For example, since the operating states andthe environmental conditions vary greatly in automotive applications,stabilizing compressed self-ignition is a major problem. The combustionby compressed self-ignition has not been put to practical use for theautomobile engine yet. In order to solve the problem, for example,JP4,082,292B2 proposes that an ignition plug ignites the mixture gas,when compressed self-ignition hardly occurs because of a lowcombustion-chamber temperature. By igniting the mixture gas immediatelybefore the compression top dead center, the pressure around the ignitionplug increases to facilitate the compressed self-ignition.

Unlike the technology disclosed in JP4,082,292B2 in which the compressedself-ignition is assisted by the ignition of the ignition plug, thepresent applicant rather proposes SPCCI (SPark Controlled CompressionIgnition) combustion which is a combination of SI (Spark Ignition)combustion and CI (Compression Ignition) combustion. The SI combustionis combustion accompanied by the flame propagation initiated by forciblyigniting the mixture gas inside the combustion chamber. The CIcombustion is combustion initiated by the mixture gas inside thecombustion chamber carrying out the compressed self-ignition. The SPCCIcombustion is combustion in which, when the mixture gas inside thecombustion chamber is forcibly ignited to start the combustion by flamepropagation, the unburnt mixture gas inside the combustion chambercombusts by the compression-ignition because of a pressure buildup dueto the heat generation and the flame propagation of the SI combustion.Since the SPCCI combustion includes the CI combustion, it is one form of“the combustion by compression-ignition.”

The CI combustion takes place, when the in-cylinder temperature reachesan ignition temperature defined by the composition of the mixture gas.Fuel efficiency can be maximized, if the in-cylinder temperature reachesthe ignition temperature near a compression top dead center and the CIcombustion takes place. The in-cylinder temperature increases accordingto the increase in the in-cylinder pressure. The in-cylinder pressure inthe SPCCI combustion is a result of two pressure buildups of a pressurebuildup by the compression work of the piston in a compression stroke,and a pressure buildup caused by the heat generation of the SIcombustion.

Here, if the CI combustion takes place near a compression top deadcenter because of a high in-cylinder temperature at a compressionstarting timing due to a high ambient temperature, etc., the in-cylinderpressure excessively increases to create excessive combustion noise. Inthis case, combustion noise can be reduced if the ignition timing isretarded. However, if the ignition timing is retarded, since the CIcombustion takes place when the piston falls considerably in theexpansion stroke, fuel efficiency is decreased. Since the pressurebuildup caused by the heat generation of the SI combustion can beutilized in the SPCCI combustion, for example, it is effective to lowerthe effective compression ratio and to reduce the pressure buildup bythe compression work of the piston in order to achieve both thereduction of combustion noise and improvement in fuel efficiency. Thus,combustion noise can be kept suitable, without decreasing fuelefficiency.

In order to put to practical use the engine which performs the SPCCIcombustion, it is necessary to take into consideration other controlfactors relevant to the in-cylinder temperature, other than effectivecompression ratio. However, since the SPCCI combustion is a newcombustion system, no one has found other control factors until now.

Since the SPCCI combustion is compression-ignition combustion,combustion noise tends to be increased. The present inventors found thatit was necessary to adjust the temperature inside the combustion chamberat the start timing of the CI combustion to a suitable temperature, inorder to achieve a stable SPCCI combustion, while reducing combustionnoise. If the temperature inside the combustion chamber is low, theignitability of the CI combustion falls. Combustion noise increases asthe temperature inside the combustion chamber goes up.

The temperature inside the combustion chamber mainly depends on thegeometric compression ratio of the engine, and the temperature and/oramount of gas introduced into the combustion chamber. Properties relatedto the gas introduced into the combustion chamber depend on thesupercharging pressure of the supercharger, and the valve timing of theintake valve, especially when boosting. The optimal superchargingpressure was unknown so far.

SUMMARY OF THE DISCLOSURE

The present inventors succeeded in finding a relation between thesupercharging pressure and the close timing of the intake valve, whichappropriately causes SPCCI combustion, as a result of repeated anddiligent examinations of SPCCI combustion. The present inventors came toinvent a control device of a compression-ignition engine based on thisknowledge.

Specifically, the technology disclosed herein relates to a controldevice of a compression-ignition engine.

The engine includes a fuel injection part configured to inject fuel tobe supplied in a combustion chamber, a variable valve operatingmechanism configured to change a valve timing of an intake valve, anignition part configured to ignite a mixture gas inside the combustionchamber, a supercharger configured to boost gas introduced into thecombustion chamber, a measurement part configured to measure a parameterrelated to an operating state of the engine, and a control partconfigured to perform a calculation according to a control logiccorresponding to the operating state of the engine, in response to themeasurement of the measurement part, and output a signal to the fuelinjection part, the variable valve operating mechanism, the ignitionpart, and the supercharger.

The control part outputs the signal to the fuel injection part, thevariable valve operating mechanism and the supercharger so that agas-fuel ratio (G/F) that is a weight ratio of the entire gas of themixture gas inside the combustion chamber to the fuel becomes leanerthan a stoichiometric air fuel ratio, and an air-fuel ratio (A/F) thatis a weight ratio of air contained in the mixture gas to the fuelbecomes the stoichiometric air fuel ratio or richer than thestoichiometric air fuel ratio, while causing the supercharger to boost,and outputs the signal to the ignition part so that the unburnt mixturegas combusts by self-ignition after the ignition part ignites themixture gas inside the combustion chamber.

The control part determines a supercharging pressure P by thesupercharger, and determines a close timing IVC of the intake valve. Thecontrol part determines, according to the control logic, the closetiming IVC (deg.aBDC) so that the supercharging pressure P (kPa)satisfies the following expression.P≥8.0×10⁻¹¹IVC⁶−1.0×10⁻⁸IVC⁵3.0×10⁻⁷IVC⁴−4.0×10⁻⁶IVC³+0.0068IVC²−0.3209IVC+116.63  (A)

The ignition part ignites the mixture gas inside the combustion chamberin response to the signal from the control part. The combustion startsby flame propagation and then the unburnt mixture gas combusts byself-ignition to complete the combustion. That is, this engine performsthe SPCCI (SPark Controlled Compression Ignition) combustion.

With this engine, the G/F of the mixture gas is made leaner than thestoichiometric air fuel ratio and the A/F is made to be thestoichiometric air fuel ratio or richer than the stoichiometric air fuelratio. By making the G/F lean, fuel efficiency of the engine improves,and by making the A/F the stoichiometric air fuel ratio, emissionperformance improves by using a catalyst device.

In addition, the boost by the supercharger results in keeping the G/F ofthe mixture gas leaner than the stoichiometric air fuel ratio. This isadvantageous in improving the fuel efficiency of the engine.

When controlling the engine, the control part first determines thesupercharging pressure P by the supercharger. When the superchargingpressure P is set, the control part determines the close timing IVC ofthe intake valve so that the expression (A) is satisfied. By setting theclose timing so as to satisfy the expression (A), the engine can performa stable SPCCI combustion which is a combination of SI (Spark Ignition)combustion and CI (Compression Ignition) combustion while keepingcombustion noise within the allowable range and the G/F leanenvironment, even under various conditions with different situations ofthe combustion chamber.

The index for determining the close timing IVC of the intake valve,which is available when controlling the engine for performing the SPCCIcombustion, has been unknown until now.

The controlling method defines the relationship between thesupercharging pressure P and the close timing IVC of the intake valve inorder to achieve a suitable SPCCI combustion. When operating the engine,the control part can set the close timing IVC of the intake valve withinthe range in which the relationship is satisfied. Accordingly, theengine for performing the SPCCI combustion can put to practical use.

The supercharging pressure (kPa) may be determined so as to satisfy thefollowing expression.P≥150   (B)

This is advantageous in using a comparatively small sized supercharger.

The control part may determine a target supercharging pressurecorresponding to the operating state of the engine. The control part maydetermine the close timing IVC (deg.aBDC) so that, when the targetsupercharging pressure is used as the supercharging pressure P (kPa),the supercharging pressure P (kPa) satisfies the expression.

For example, when the engine load increases and the fuel amount suppliedinto the combustion chamber increases, by performing the boost by thesupercharger, the G/F of the mixture gas is leaner than thestoichiometric air fuel ratio and the A/F is the stoichiometric air fuelratio or richer than the stoichiometric air fuel ratio, which isadvantageous in improving fuel efficiency.

The close timing IVC of the intake valve may change as the operatingstate of the engine changes, and the close timing IVC (deg.aBDC) may bedetermined for each operating state so that one of the expressions (A)and (B) is satisfied.

Thus, the engine can stably perform the SPCCI combustion in variousoperating states.

The engine may operate in a high-load operating state at a given load orhigher.

In general CI combustion, since the pressure fluctuation at the ignitionis relatively large, the combustion noise becomes too large when theengine load is high, and may exceed the allowable range. In this regard,in the SPCCI combustion, the SI combustion is performed at the start ofthe combustion and the pressure fluctuation at the ignition in the SIcombustion is small. Thus, combustion noise can be reduced even when theengine load is high.

The engine may operate in a maximum load operating state. That is theengine may perform the SPCCI combustion in the maximum load operatingstate.

A geometric compression ratio ε of the engine may be set so as tosatisfy 10≤ε<21. In this way, the geometric compression ratio can be setsuitably.

The engine may be provided with an exhaust gas recirculation (EGR)system configured to introduce exhaust gas into the combustion chamber.The control part may output the signal to the EGR system and theignition part so that a heat amount ratio used as an index related to aratio of an amount of heat generated when the mixture gas inside thecombustion chamber combusts by flame propagation to the entire amount ofheat generated when the mixture gas combusts, becomes a target heatamount ratio defined corresponding to the operating state of the engine.

The heat amount ratio of the SPCCI combustion is less than 100%. Theheat amount ratio of the combustion mode where the combustion completesonly by flame propagation without the combustion by compression ignition(i.e., SI combustion) is 100%.

If the heat amount ratio is increased in the SPCCI combustion, the ratioof the SI combustion increases, which is advantageous in reducingcombustion noise. Whereas, if the heat amount ratio is lowered in theSPCCI combustion, the ratio of the CI combustion increases, which isadvantageous in improving fuel efficiency. The heat amount ratio changesby changing the temperature of the combustion chamber and/or theignition timing. For example, when the temperature inside the combustionchamber is high, the CI combustion starts at an early timing, and theheat amount ratio becomes low. Further, when the ignition timing isadvanced, the SI combustion starts at an early timing, and the heatamount ratio becomes high. By the control part outputting signals to theEGR system and the ignition part so that the heat amount ratio becomesthe target heat amount ratio defined in accordance with the operatingstate of the engine, both the reduction of combustion noise and theimprovement of fuel efficiency can be achieved.

The control part may output a signal to the EGR system and the ignitionpart so that the heat amount ratio becomes higher when the load of theengine is higher.

When the engine load increases, the amount of fuel supplied into thecombustion chamber increases and the temperature inside the combustionchamber becomes high. By increasing the heat amount ratio of the SPCCIcombustion when the engine load is high, the combustion noise isreduced.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a view illustrating a configuration of an engine.

FIG. 2 is a view illustrating a configuration of a combustion chamber,where an upper figure corresponds to a plan view of the combustionchamber, and a lower figure is a cross-sectional view taken along a lineII-II.

FIG. 3 is a plan view illustrating a configuration of the combustionchamber and an intake system.

FIG. 4 is a block diagram illustrating a configuration of an enginecontrol device.

FIG. 5 is a graph illustrating a waveform of SPCCI combustion.

FIG. 6 illustrates maps of the engine, where an upper figure is a mapwhen the engine is warm, a middle figure is a map when the engine ishalf warm, and a lower figure is a map when the engine is cold.

FIG. 7 illustrates the details of the map when the engine is warm.

FIG. 8 illustrates charts of a fuel injection timing, an ignitiontiming, and a combustion waveform in each operating range of the map ofFIG. 7.

FIG. 9 illustrates a layer structure of the engine map.

FIG. 10 is a flowchart illustrating a control process according to alayer selection of the map.

An upper figure of FIG. 11 is a graph illustrating a relation between anengine load and an open timing of an intake valve in Layer 2, and alower figure thereof is a graph illustrating a relation between anengine speed and the open timing of the intake valve in Layer 2.

An upper figure of FIG. 12 is a graph illustrating a relation betweenthe engine load and the open timing of the intake valve in Layer 3, amiddle figure thereof is a graph illustrating a relation between theengine load and a close timing of an exhaust valve in Layer 3, and alower figure thereof is a graph illustrating a relation between theengine load, and an overlap period of the intake valve and the exhaustvalve in Layer 3.

FIG. 13 is a flowchart illustrating a process of an operation control ofthe engine executed by an ECU.

FIG. 14 illustrates a relation between the engine load and a target SIratio.

FIG. 15 is a graph illustrating an occurring range of the SPCCIcombustion versus an exhaust gas recirculation (EGR) rate in Layer 2.

FIG. 16 is one example of a matrix image utilized in order to determinea relation between a geometric compression ratio and a close timing ofthe intake valve where the SPCCI combustion is possible in Layer 2.

An upper figure of FIG. 17 illustrates a relation between the geometriccompression ratio and the close timing of the intake valve where theSPCCI combustion is possible in Layer 2 when high octane fuel is used,and a lower figure thereof illustrates a relation between the geometriccompression ratio and the close timing of the intake valve where theSPCCI combustion is possible in Layer 2 when low octane fuel is used.

FIG. 18 is a graph illustrating a range where the SPCCI combustion isstabilized versus a gas-fuel ratio (G/F) in Layer 3.

FIG. 19 is one example of a matrix image utilized in order to determinethe relation between the geometric compression ratio and the closetiming of the intake valve where the SPCCI combustion is possible inLayer 3.

An upper figure of FIG. 20 illustrates a relation between the geometriccompression ratio and the close timing of the intake valve where theSPCCI combustion is possible in Layer 3 when high octane fuel is used,and a lower figure thereof illustrates a relation between the geometriccompression ratio and the close timing of the intake valve where theSPCCI combustion is possible in Layer 3 when low octane fuel is used.

An upper figure of FIG. 21 illustrates a relation between the geometriccompression ratio and the close timing of the intake valve where theSPCCI combustion is possible in Layer 2 and Layer 3 when high octanefuel is used, and a lower figure thereof illustrates a relation betweenthe geometric compression ratio and the close timing of the intake valvewhere the SPCCI combustion is possible in Layer 2 and Layer 3 when lowoctane fuel is used.

FIG. 22 is a flowchart illustrating a procedure of controlling acompression-ignition engine.

An upper figure of FIG. 23 illustrates a relation between asupercharging pressure and the close timing of the intake valve where aG/F lean SPCCI combustion is possible in Layer 2 when the geometriccompression ratio is set relatively high, and a lower figure thereofillustrates a relation between a supercharging pressure and the closetiming of the intake valve where the G/F lean SPCCI combustion ispossible in Layer 2 when the geometric compression ratio is setrelatively low.

FIG. 24 is a flowchart illustrating a procedure of controlling thecompression-ignition engine.

DETAILED DESCRIPTION OF THE DISCLOSURE

Hereinafter, one embodiment of a method of implementing control logic ofa compression-ignition engine will be described in detail with referenceto the accompanying drawings. The following description is one exampleof the engine and the method of implementing the control logic.

FIG. 1 is a view illustrating a configuration of thecompression-ignition engine. FIG. 2 is a view illustrating aconfiguration of a combustion chamber of the engine. FIG. 3 is a viewillustrating a configuration of the combustion chamber and an intakesystem. Note that in FIG. 1, an intake side is the left side in thedrawing, and an exhaust side is the right side in the drawing. In FIGS.2 and 3, the intake side is the right side in the drawings, and theexhaust side is the left side in the drawings. FIG. 4 is a block diagramillustrating a configuration of a control device of the engine.

An engine 1 is a four-stroke engine which operates by a combustionchamber 17 repeating an intake stroke, a compression stroke, anexpansion stroke, and an exhaust stroke. The engine 1 is mounted on anautomobile with four wheels. The automobile travels by operating theengine 1. Fuel of the engine 1 is gasoline in this example. The fuel maybe a liquid fuel containing at least gasoline. The fuel may be gasolinecontaining, for example, bioethanol.

(Engine Configuration)

The engine 1 includes a cylinder block 12 and a cylinder head 13 placedthereon. A plurality of cylinders 11 are formed inside the cylinderblock 12. In FIGS. 1 and 2, only one cylinder 11 is illustrated. Theengine 1 is a multi-cylinder engine.

A piston 3 is slidably inserted in each cylinder 11. The pistons 3 areconnected with a crankshaft 15 through respective connecting rods 14.Each piston 3 defines the combustion chamber 17, together with thecylinder 11 and the cylinder head 13. Note that the term “combustionchamber” may be used in a broad sense. That is, the term “combustionchamber” may refer to a space formed by the piston 3, the cylinder 11,and the cylinder head 13, regardless of the position of the piston 3.

As illustrated in the lower figure of FIG. 2, a lower surface of thecylinder head 13, i.e., a ceiling surface of the combustion chamber 17,is comprised of a slope 1311 and a slope 1312. The slope 1311 is arising gradient from the intake side toward an injection axial center X2of an injector 6 which will be described later. The slope 1312 is arising gradient from the exhaust side toward the injection axial centerX2. The ceiling surface of the combustion chamber 17 is a so-called“pent-roof” shape.

An upper surface of the piston 3 is bulged toward the ceiling surface ofthe combustion chamber 17. A cavity 31 is formed in the upper surface ofthe piston 3. The cavity 31 is a dent in the upper surface of the piston3. The cavity 31 has a shallow pan shape in this example. The center ofthe cavity 31 is offset at the exhaust side with respect to a centeraxis X1 of the cylinder 11.

A geometric compression ratio ε of the engine 1 is set so as to satisfy1≤ε<30, and preferably satisfy 10≤ε<21. The engine 1 which will bedescribed later performs SPCCI (SPark Controlled Compression Ignition)combustion that is a combination of SI (Spark Ignition) combustion andCI (Compression Ignition) combustion in a part of operating ranges. TheSPCCI combustion controls the CI combustion using a heat generation anda pressure buildup by the SI combustion. The engine 1 is thecompression-ignition engine. However, in this engine 1, temperature ofthe combustion chamber 17, when the piston 3 is at a compression topdead center (i.e., compression end temperature), does not need to beincreased. In the engine 1, the geometric compression ratio can be setcomparatively low. The low geometric compression ratio becomesadvantageous in reduction of cooling loss and mechanical loss. Forengines using regular gasoline (a low octane fuel of which octane numberis about 91), the geometric compression ratio of the engine 1 is 14-17,and for those using high octane gasoline (high octane fuel of whichoctane number is about 96), the geometric compression ratio is 15-18.

An intake port 18 is formed in the cylinder head 13 for each cylinder11. As illustrated in FIG. 3, each intake port 18 has a first intakeport 181 and a second intake port 182. The intake port 18 communicateswith the corresponding combustion chamber 17. Although the detailedillustration of the intake port 18 is omitted, it is a so-called “tumbleport.” That is, the intake port 18 has such a shape that a tumble flowis formed in the combustion chamber 17.

Each intake valve 21 is disposed in the intake ports 181 and 182. Theintake valve 21 opens and closes a channel between the combustionchamber 17 and the intake port 181 or 182. The intake valves 21 areopened and closed at given timings by a valve operating mechanism. Thevalve operating mechanism may be a variable valve operating mechanismwhich varies the valve timing and/or valve lift. In this example, asillustrated in FIG. 4, the variable valve operating mechanism has anintake-side electric S-VT (Sequential-Valve Timing) 23. The intake-sideelectric S-VT 23 continuously varies a rotation phase of an intake camshaft within a given angle range. The open timing and the close timingof the intake valve 21 vary continuously. Note that the electric S-VTmay be replaced with a hydraulic S-VT, as the intake valve operatingmechanism.

An exhaust port 19 is also formed in the cylinder head 13 for eachcylinder 11. As illustrated in FIG. 3, each exhaust port 19 also has afirst exhaust port 191 and a second exhaust port 192. The exhaust port19 communicates with the corresponding combustion chamber 17.

Each exhaust valve 22 is disposed in the exhaust ports 191 and 192. Theexhaust valve 22 opens and closes a channel between the combustionchamber 17 and the exhaust port 191 or 192. The exhaust valves 22 areopened and closed at a given timing by a valve operating mechanism. Thevalve operating mechanism may be a variable valve operating mechanismwhich varies the valve timing and/or valve lift. In this example, asillustrated in FIG. 4, the variable valve operating mechanism has anexhaust-side electric SVT 24. The exhaust-side electric S-VT 24continuously varies a rotation phase of an exhaust cam shaft within agiven angle range. The open timing and the close timing of the exhaustvalve 22 change continuously. Note that the electric S-VT may bereplaced with a hydraulic S-VT, as the exhaust valve operatingmechanism.

The intake-side electric S-VT 23 and the exhaust-side electric S-VT 24adjust the length of an overlap period where both the intake valve 21and the exhaust valve 22 are open. If the length of the overlap periodis made longer, the residual gas in the combustion chamber 17 can bepurged. Moreover, by adjusting the length of the overlap period,internal EGR (Exhaust Gas Recirculation) gas can be introduced into thecombustion chamber 17. An internal EGR system is comprised of theintake-side electric S-VT 23 and the exhaust-side electric S-VT 24. Notethat the internal EGR system may not be comprised of the S-VT.

The injector 6 is attached to the cylinder head 13 for each cylinder 11.Each injector 6 directly injects fuel into the combustion chamber 17.The injector 6 is one example of a fuel injection part. The injector 6is disposed in a valley part of the pent roof where the slope 1311 andthe slope 1312 meet. As illustrated in FIG. 2, the injection axialcenter X2 of the injector 6 is located at the exhaust side of the centeraxis X1 of the cylinder 11. The injection axial center X2 of theinjector 6 is parallel to the center axis X1. The injection axial centerX2 of the injector 6 and the center of the cavity 31 are in agreementwith each other. The injector 6 faces the cavity 31. Note that theinjection axial center X2 of the injector 6 may be in agreement with thecenter axis X1 of the cylinder 11. In such a configuration, theinjection axial center X2 of the injector 6 and the center of the cavity31 may be in agreement with each other.

Although detailed illustration is omitted, the injector 6 is comprisedof a multi nozzle-port type fuel injection valve having a plurality ofnozzle ports. As illustrated by two-dot chain lines in FIG. 2, theinjector 6 injects fuel so that the fuel spreads radially from thecenter of the combustion chamber 17. The injector 6 has ten nozzle portsin this example, and the nozzle port is disposed so as to be equallyspaced in the circumferential direction.

The injectors 6 are connected to a fuel supply system 61. The fuelsupply system 61 includes a fuel tank 63 configured to store fuel, and afuel supply passage 62 which connects the fuel tank 63 to the injector6. In the fuel supply passage 62, a fuel pump 65 and a common rail 64are provided. The fuel pump 65 pumps fuel to the common rail 64. Thefuel pump 65 is a plunger pump driven by the crankshaft 15 in thisexample. The common rail 64 stores fuel pumped from the fuel pump 65 ata high fuel pressure. When the injector 6 is opened, the fuel stored inthe common rail 64 is injected into the combustion chamber 17 from thenozzle ports of the injector 6. The fuel supply system 61 can supplyfuel to the injectors 6 at a high pressure of 30 MPa or higher. Thepressure of fuel supplied to the injector 6 may be changed according tothe operating state of the engine 1. Note that the configuration of thefuel supply system 61 is not limited to the configuration describedabove.

An ignition plug 25 is attached to the cylinder head 13 for eachcylinder 11. The ignition plug 25 forcibly ignites a mixture gas insidethe combustion chamber 17. The ignition plug 25 is disposed at theintake side of the center axis X1 of the cylinder 11 in this example.The ignition plug 25 is located between the two intake ports 181 and 182of each cylinder. The ignition plug 25 is attached to the cylinder head13 so as to incline downwardly toward the center of the combustionchamber 17. As illustrated in FIG. 2, the electrode of the ignition plug25 faces to the inside of the combustion chamber 17 and is located nearthe ceiling surface of the combustion chamber 17. Note that the ignitionplug 25 may be disposed at the exhaust side of the center axis X1 of thecylinder 11. Moreover, the ignition plug 25 may be disposed on thecenter axis X1 of the cylinder 11.

An intake passage 40 is connected to one side surface of the engine 1.The intake passage 40 communicates with the intake port 18 of eachcylinder 11. Gas introduced into the combustion chamber 17 flows throughthe intake passage 40. An air cleaner 41 is disposed in an upstream endpart of the intake passage 40. The air cleaner 41 filters fresh air. Asurge tank 42 is disposed near the downstream end of the intake passage40. Part of the intake passage 40 downstream of the surge tank 42constitutes independent passages branched from the intake passage 40 foreach cylinder 11. The downstream end of each independent passage isconnected to the intake port 18 of each cylinder 11.

A throttle valve 43 is disposed between the air cleaner 41 and the surgetank 42 in the intake passage 40. The throttle valve 43 adjusts anintroducing amount of the fresh air into the combustion chamber 17 byadjusting an opening of the throttle valve.

A supercharger 44 is also disposed in the intake passage 40, downstreamof the throttle valve 43. The supercharger 44 boosts gas to beintroduced into the combustion chamber 17. In this example, thesupercharger 44 is a mechanical supercharger driven by the engine 1. Themechanical supercharger 44 may be a root, Lysholm, vane, or acentrifugal type.

An electromagnetic clutch 45 is provided between the supercharger 44 andthe engine 1. The electromagnetic clutch 45 transmits a driving forcefrom the engine 1 to the supercharger 44 or disengages the transmissionof the driving force between the supercharger 44 and the engine 1. Aswill be described later, an ECU 10 switches the disengagement andengagement of the electromagnetic clutch 45 to switch the supercharger44 between ON and OFF.

An intercooler 46 is disposed downstream of the supercharger 44 in theintake passage 40. The intercooler 46 cools gas compressed by thesupercharger 44. The intercooler 46 may be of a water cooling type or anoil cooling type, for example.

A bypass passage 47 is connected to the intake passage 40. The bypasspassage 47 connects an upstream part of the supercharger 44 to adownstream part of the inter cooler 46 in the intake passage 40 so as tobypass the supercharger 44 and the inter cooler 46. An air bypass valve48 is disposed in the bypass passage 47. The air bypass valve 48 adjustsa flow rate of gas flowing in the bypass passage 47.

The ECU 10 fully opens the air bypass valve 48 when the supercharger 44is turned OFF (i.e., when the electromagnetic clutch 45 is disengaged).The gas flowing through the intake passage 40 bypasses the supercharger44 and is introduced into the combustion chamber 17 of the engine 1. Theengine 1 operates in a non-supercharged state, i.e., a naturalaspiration state.

When the supercharger 44 is turned ON, the engine 1 operates in asupercharged state. The ECU 10 adjusts an opening of the air bypassvalve 48 when the supercharger 44 is turned ON (i.e., when theelectromagnetic clutch 45 is engaged). A portion of the gas which passedthrough the supercharger 44 flows back toward upstream of thesupercharger 44 through the bypass passage 47. When the ECU 10 adjuststhe opening of the air bypass valve 48, a supercharging pressure of gasintroduced into the combustion chamber 17 changes. Note that the term“supercharging” as used herein refers to a situation where the pressureinside the surge tank 42 exceeds an atmospheric pressure, and“non-supercharging” refers to a situation where the pressure inside thesurge tank 42 becomes below the atmospheric pressure.

In this example, a supercharging system 49 is comprised of thesupercharger 44, the bypass passage 47, and the air bypass valve 48.

The engine 1 has a swirl generating part which generates a swirl flowinside the combustion chamber 17. As illustrated in FIG. 3, the swirlgenerating part has a swirl control valve 56 attached to the intakepassage 40. Among a primary passage 401 coupled to the first intake port181 and a secondary passage 402 coupled to the second intake port 182,the swirl control valve 56 is disposed in the secondary passage 402. Theswirl control valve 56 is an opening control valve which is configuredto choke a cross section of the secondary passage 402. When the openingof the swirl control valve 56 is small, since an intake flow rate of airflowing into the combustion chamber 17 from the first intake port 181 isrelatively large, and an intake flow rate of air flowing into thecombustion chamber 17 from the second intake port 182 is relativelysmall, the swirl flow inside the combustion chamber 17 becomes stronger.On the other hand, when the opening of the swirl control valve 56 islarge, since the intake flow rates of air flowing into the combustionchamber 17 from the first intake port 181 and the second intake port 182become substantially equal, the swirl flow inside the combustion chamber17 becomes weaker. When the swirl control valve 56 is fully opened, theswirl flow will not occur. Note that the swirl flow circulatescounterclockwise in FIG. 3, as illustrated by white arrows (also seewhite arrows in FIG. 2).

An exhaust passage 50 is connected to the other side surface of theengine 1. The exhaust passage 50 communicates with the exhaust port 19of each cylinder 11. The exhaust passage 50 is a passage through whichexhaust gas discharged from the combustion chambers 17 flows. Althoughdetailed illustration is omitted, an upstream part of the exhaustpassage 50 constitutes independent passages branched from the exhaustpassage 50 for each cylinder 11. The upper end of the independentpassage is connected to the exhaust port 19 of each cylinder 11.

An exhaust gas purification system having a plurality of catalyticconverters is disposed in the exhaust passage 50. Although illustrationis omitted, an upstream catalytic converter is disposed inside an engineroom. The upstream catalytic converter has a three-way catalyst 511 anda GPF (Gasoline Particulate Filter) 512. The downstream catalyticconverter is disposed outside the engine room. The downstream catalyticconverter has a three-way catalyst 513. Note that the exhaust gaspurification system is not limited to the illustrated configuration. Forexample, the GPF may be omitted. Moreover, the catalytic converter isnot limited to those having the three-way catalyst. Further, the orderof the three-way catalyst and the GPF may suitably be changed.

Between the intake passage 40 and the exhaust passage 50, an EGR passage52 which constitutes an external EGR system is connected. The EGRpassage 52 is a passage for recirculating part of the exhaust gas to theintake passage 40. The upstream end of the EGR passage 52 is connectedbetween the upstream catalytic converter and the downstream catalyticconverter in the exhaust passage 50. The downstream end of the EGRpassage 52 is connected to an upstream part of the supercharger 44 inthe intake passage 40. EGR gas flowing through the EGR passage 52 flowsinto the upstream part of the supercharger 44 in the intake passage 40,without passing through the air bypass valve 48 of the bypass passage47.

An EGR cooler 53 of a water cooling type is disposed in the EGR passage52. The EGR cooler 53 cools the exhaust gas. An EGR valve 54 is alsodisposed in the EGR passage 52. The EGR valve 54 adjusts a flow rate ofthe exhaust gas flowing through the EGR passage 52. By adjusting theopening of the EGR valve 54, an amount of the cooled exhaust gas, i.e.,a recirculating amount of external EGR gas can be adjusted.

In this example, an EGR system 55 is comprised of the external EGRsystem and the internal EGR system. The external EGR system can supplythe exhaust gas to the combustion chamber 17 that is lower intemperature than the internal EGR system.

The control device of the compression-ignition engine includes the ECU(Engine Control Unit) 10 for operating the engine 1. The ECU 10 is acontroller based on a well-known microcomputer, and as illustrated inFIG. 4, includes a central processing unit (CPU) 101 which executes acomputer program, memory 102 which, for example, is comprised of a RAM(Random Access Memory) and/or a ROM (Read Only Memory), and stores theprogram and data, and an input/output bus 103 which inputs and outputsan electrical signal. The ECU 10 is one example of the control part.

As illustrated in FIGS. 1 and 4, various kinds of sensors SW1-SW17 areconnected to the ECU 10. The sensors SW1-SW17 output signals to the ECU10. The sensors include the following sensors:

Airflow sensor SW1: Disposed downstream of the air cleaner 41 in theintake passage 40, and measures a flow rate of fresh air flowing throughthe intake passage 40;

First intake-air temperature sensor SW2: Disposed downstream of the aircleaner 41 in the intake passage 40, and measures the temperature offresh air flowing through the intake passage 40;

First pressure sensor SW3: Disposed downstream of the connected positionof the EGR passage 52 in the intake passage 40 and upstream of thesupercharger 44, and measures the pressure of gas flowing into thesupercharger 44;

Second intake-air temperature sensor SW4: Disposed downstream of thesupercharger 44 in the intake passage 40 and upstream of the connectedposition of the bypass passage 47, and measures the temperature of gasflowed out of the supercharger 44;

Second pressure sensor SW5: Attached to the surge tank 42, and measuresthe pressure of gas downstream of the supercharger 44;

Pressure indicating sensors SW6: Attached to the cylinder head 13corresponding to each cylinder 11, and measures the pressure inside eachcombustion chamber 17;

Exhaust temperature sensor SW7: Disposed in the exhaust passage 50, andmeasures the temperature of the exhaust gas discharged from thecombustion chamber 17;

Linear O₂ sensor SW8: Disposed upstream of the upstream catalyticconverter in the exhaust passage 50, and measures the oxygenconcentration of the exhaust gas;

Lambda O₂ sensor SW9: Disposed downstream of the three-way catalyst 511in the upstream catalytic converter, and measures the oxygenconcentration of the exhaust gas;

Water temperature sensor SW10: Attached to the engine 1 and measures thetemperature of coolant;

Crank angle sensor SW11: Attached to the engine 1 and measures therotation angle of the crankshaft 15;

Accelerator opening sensor SW12: Attached to an accelerator pedalmechanism and measures the accelerator opening corresponding to anoperating amount of the accelerator pedal;

Intake cam angle sensor SW13: Attached to the engine 1 and measures therotation angle of an intake cam shaft;

Exhaust cam angle sensor SW14: Attached to the engine 1 and measures therotation angle of an exhaust cam shaft;

EGR pressure difference sensor SW15: Disposed in the EGR passage 52 andmeasures a pressure difference between the upstream and the downstreamof the EGR valve 54;

Fuel pressure sensor SW16: Attached to the common rail 64 of the fuelsupply system 61, and measures the pressure of fuel supplied to theinjector 6; and

Third intake-air temperature sensor SW17: Attached to the surge tank 42,and measures the temperature of gas inside the surge tank 42, i.e., thetemperature of intake air introduced into the combustion chamber 17.

The ECU 10 determines the operating state of the engine 1 based on thesignals of the sensors SW1-SW17, and calculates a control amount of eachdevice according to the control logic defined beforehand. The controllogic is stored in the memory 102. The control logic includescalculating a target amount and/or the control amount by using a mapstored in the memory 102.

The ECU 10 outputs electrical signals according to the calculatedcontrol amounts to the injectors 6, the ignition plugs 25, theintake-side electric S-VT 23, the exhaust-side electric S-VT 24, thefuel supply system 61, the throttle valve 43, the EGR valve 54, theelectromagnetic clutch 45 of the supercharger 44, the air bypass valve48, and the swirl control valve 56.

For example, the ECU 10 sets a target torque of the engine 1 based onthe signal of the accelerator opening sensor SW12 and the map, anddetermines a target supercharging pressure. The ECU 10 then performs afeedback control for adjusting the opening of the air bypass valve 48based on the target supercharging pressure and the pressure differencebefore and after the supercharger 44 obtained from the signals of thefirst pressure sensor SW3 and the second pressure sensor SW5 so that thesupercharging pressure becomes the target supercharging pressure.

Moreover, the ECU 10 sets a target EGR ratio (i.e., a ratio of the EGRgas to the entire gas inside the combustion chamber 17) based on theoperating state of the engine 1 and the map. The ECU 10 then determinesa target EGR gas amount based on the target EGR ratio and an intake airamount based on the signal of the accelerator opening sensor SW12, andperforms a feedback control for adjusting the opening of the EGR valve54 based on the pressure difference before and after the EGR valve 54obtained from the signal of the EGR pressure difference sensor SW15 sothat the external EGR gas amount introduced into the combustion chamber17 becomes the target EGR gas amount.

Further, the ECU 10 performs an air-fuel ratio feedback control when agiven control condition is satisfied. For example, the ECU 10 adjuststhe fuel injection amount of the injector 6 based on the oxygenconcentration of the exhaust gas which is measured by the linear O₂sensor SW8 and the lambda O₂ sensor SW9 so that the air-fuel ratio ofthe mixture gas becomes a desired value.

Note that the details of other controls of the engine 1 executed by theECU 10 will be described later.

(Concept of SPCCI Combustion)

The engine 1 performs combustion by compressed self-ignition under agiven operating state, mainly to improve fuel consumption and emissionperformance. The combustion by self-ignition varies largely at thetiming of the self-ignition, if the temperature inside the combustionchamber 17 before a compression starts is nonuniform. Thus, the engine 1performs the SPCCI combustion which is a combination of the SIcombustion and the CI combustion.

The SPCCI combustion is combustion in which the ignition plug 25forcibly ignites the mixture gas inside the combustion chamber 17 sothat the mixture gas carries out the SI combustion by flame propagation,and the temperature inside the combustion chamber 17 increases by theheat generation of the SI combustion and the pressure inside thecombustion chamber 17 increases by the flame propagation so that theunburnt mixture gas carries out the CI combustion by self-ignition.

By adjusting the heat amount of the SI combustion, the variation in thetemperature inside the combustion chamber 17 before a compression startscan be absorbed. By the ECU 10 adjusting the ignition timing, themixture gas can be self-ignited at a target timing.

In the SPCCI combustion, the heat release of the SI combustion is slowerthan the heat release in the CI combustion. As illustrated in FIG. 5,the waveform of the heat release rate of the SI combustion in the SPCCIcombustion is smaller in the rising slope than the waveform in the CIcombustion. In addition, the SI combustion is slower in the pressurefluctuation (dp/dθ) inside the combustion chamber 17 than the CIcombustion.

When the unburnt mixture gas self-ignites after the SI combustion isstarted, the waveform slope of the heat release rate may become steeper.The waveform of the heat release rate may have an inflection point X ata timing of starting the CI combustion.

After the start in the CI combustion, the SI combustion and the CIcombustion are performed in parallel. Since the CI combustion has largerheat release than the SI combustion, the heat release rate becomesrelatively large. However, since the CI combustion is performed after acompression top dead center, the waveform slope of the heat release ratedoes not become too steep. The pressure fluctuation in the CI combustion(dp/dθ) also becomes comparatively slow.

The pressure fluctuation (dp/dθ) can be used as an index representingcombustion noise. As described above, since the SPCCI combustion canreduce the pressure fluctuation (dp/dθ), it is possible to avoidexcessive combustion noise. Combustion noise of the engine 1 can be keptbelow the tolerable level.

The SPCCI combustion is completed when the CI combustion is finished.The CI combustion is shorter in the combustion period than the SIcombustion. The end timing of the SPCCI combustion becomes earlier thanthe SI combustion.

The heat release rate waveform of the SPCCI combustion is formed so thata first heat release rate part Q_(SI) formed by the SI combustion and asecond heat release rate part Q_(CI) formed by the CI combustioncontinue in this order.

Here, a SI ratio is defined as a parameter indicative of acharacteristic of the SPCCI combustion. The SI ratio is defined as anindex related to a ratio of the amount of heat generated by the SIcombustion to the entire amount of heat generated by the SPCCIcombustion. The SI ratio is a ratio of heat amount generated by the twodifferent combustion modes. If the SI ratio is high, the ratio of the SIcombustion is high, and if the SI ratio is low, the ratio in the CIcombustion is high. Alternatively, the SI ratio may be defined as aratio of the amount of heat generated by the SI combustion to the amountof heat generated by the CI combustion. That is, in a waveform 801illustrated in FIG. 5,SI ratio=Q _(SI) /Q _(CI).Here,

Q_(SI): Area of SI combustion; and

Q_(CI): Area of CI combustion.

The engine 1 generates a strong swirl flow inside the combustion chamber17 when performing the SPCCI combustion. The term “strong swirl flow”may be defined as a flow having a swirl ratio of four or higher, forexample. The swirl ratio can be defined as a value obtained by dividingan integrated value of intake flow lateral angular velocities by anengine angular velocity, where the intake flow lateral angular velocityis measured for every valve lift, and the measured values are integratedto obtain the integrated value. Although illustration is omitted, theintake flow lateral angular velocity can be obtained based onmeasurement using known rig test equipment.

When the strong swirl flow is generated in the combustion chamber 17,the swirl flow is stronger in an outer circumferential part of thecombustion chamber 17 and is relatively weaker in a central part. By thewhirlpool resulting from a velocity gradient at the boundary between thecentral part and the outer circumferential part, turbulence energybecomes higher in the central part. When the ignition plug 25 ignitesthe mixture gas in the central part, the combustion speed of the SIcombustion becomes higher by the high turbulence energy.

Flame of the SI combustion is carried by the strong swirl flow insidethe combustion chamber 17 and propagates in the circumferentialdirection. The CI combustion is performed from the outer circumferentialpart to the central part in the combustion chamber 17.

When the strong swirl flow is generated in the combustion chamber 17,the SI combustion can fully be performed before the start in the CIcombustion. Thus, the generation of combustion noise can be reduced, andthe variation in the torque between cycles can be reduced.

(Engine Operating Range)

FIGS. 6 and 7 illustrate maps according to the control of the engine 1.The maps are stored in the memory 102 of the ECU 10. The maps includethree kinds of maps, a map 501, a map 502, and a map 503. The ECU 10uses a map selected from the three kinds of maps 501, 502, and 503according to a wall temperature of the combustion chamber 17 and anintake air temperature, in order to control the engine 1. Note that thedetails of the selection of the three kinds of maps 501, 502, and 503will be described later.

The first map 501 is a map when the engine 1 is warm. The second map 502is a map when the engine 1 is half warm. The third map 503 is a map whenthe engine 1 is cold.

The maps 501, 502, and 503 are defined based on the load and the enginespeed of the engine 1. The first map 501 is roughly divided into threeareas depending on the load and the engine speed. For example, the threeareas include a low load area A1, a middle-to-high load area (A2, A3,and A4), and a high speed area A5. The low load area A1 includes idleoperation, and covers areas of a low engine speed and a middle enginespeed. The middle-to-high load area (A2, A3, and A4) are higher in theload than the low load area A1. The high speed area A5 is higher in theengine speed than the low load area A1 and the middle-to-high load area(A2, A3, and A4). The middle-to-high load area (A2, A3, and A4) isdivided into a middle load area A2, a high-load middle-speed area A3where the load is higher than the middle load area A2, and a high-loadlow-speed area A4 where the engine speed is lower than the high-loadmiddle-speed area A3.

The second map 502 is roughly divided into two areas. For example, thetwo areas include a low-to-middle speed area (B1, B2, and B3) and a highspeed area B4 where the engine speed is higher than the low-to-middlespeed area (B1, B2, and B3). The low-to-middle speed area (B1, B2, andB3) is divided into a low-to-middle load area B1 corresponding to thelow load area A1 and the middle load area A2, a high-load middle-speedarea B2, and a high-load low-speed area B3.

The third map 503 has only one area Cl, without being divided into aplurality of areas.

Here, the low speed area, the middle speed area, and the high speed areamay be defined by substantially equally dividing the entire operatingrange of the engine 1 into three areas in the engine speed direction. Inthe example of FIGS. 6 and 7, the engine speed is defined to be a lowspeed if the engine speed is lower than the engine speed N1, a highspeed if the engine speed is higher than or equal to the engine speedN2, and a middle speed if the engine speed is higher than or equal tothe engine speed N1 and lower than the engine speed N2. For example, theengine speed N1 may be about 1,200 rpm, and the engine speed N2 may beabout 4,000 rpm.

Moreover, the low load area may be an area including an operating statewith the light load, the high load area may be an area including anoperating state with full load, and the middle load area may be an areabetween the low load area and the high load area. Moreover, the low loadarea, the middle load area, and the high load area may be defined bysubstantially equally dividing the entire operating range of the engine1 into three areas in the load direction.

The maps 501, 502, and 503 in FIG. 6 illustrate the states andcombustion modes of the mixture gas in the respective areas. A map 504in FIG. 7 corresponds to the first map 501, and illustrates the stateand combustion mode of the mixture gas in each area of the map, theopening of the swirl control valve 56 in each area, and a driving areaand a non-driving area of the supercharger 44. The engine 1 performs theSPCCI combustion in the low load area A1, the middle load area A2, thehigh-load middle-speed area A3, the high-load low-speed area A4, thelow-to-middle load area B1, the high-load middle-speed area B2, and thehigh-load low-speed area B3. The engine 1 performs the SI combustion inother areas, specifically, in the high speed area A5, the high speedarea B4, and the area C1.

(Operation of Engine in Each Area)

Below, the operation of the engine 1 in each area of the map 504 in FIG.7 will be described in detail with reference to the fuel injectiontiming and the ignition timing which are illustrated in FIG. 8. Thehorizontal axis in FIG. 8 is a crank angle. Note that the referencenumerals 601, 602, 603, 604, 605, and 606 in FIG. 8 correspond to theoperating states of the engine 1 indicated by the reference numerals601, 602, 603, 604, 605, and 606 in the map 504 of FIG. 7, respectively.

(Low load Area)

The engine 1 performs the SPCCI combustion when the engine 1 operates inthe low load area A1.

The reference numeral 601 in FIG. 8 indicates fuel injection timings(reference numerals 6011 and 6012), an ignition timing (referencenumeral 6013), and a combustion waveform (i.e., a waveform indicating achange in the heat release rate with respect to the crank angle:reference numeral 6014), when the engine 1 operates in the operatingstate 601 in the low load area A1. The reference numeral 602 indicatesfuel injection timings (reference numerals 6021 and 6022), an ignitiontiming (reference numeral 6023), and a combustion waveform (referencenumeral 6024), when the engine 1 operates in the operating state 602 inthe low load area A1. The reference numeral 603 indicates fuel injectiontimings (reference numerals 6031 and 6032), an ignition timing(reference numeral 6033), and a combustion waveform (reference numeral6034), when the engine 1 operates in the operating state 603 in the lowload area A1. The operating states 601, 602, and 603 have the sameengine speed, but different loads. The operating state 601 has thelowest load (i.e., light load), the operating state 602 has the secondlowest load (i.e., low load), and the operating state 603 has themaximum load among these states.

In order to improve the fuel efficiency of the engine 1, the EGR system55 introduces the EGR gas into the combustion chamber 17. For example,the intake-side electric S-VT 23 and the exhaust-side electric S-VT 24are provided with a positive overlap period where both the intake valve21 and the exhaust valve 22 are opened near an exhaust top dead center.Part of the exhaust gas discharged from the combustion chamber 17 intothe intake port 18 and the exhaust port 19 is re-introduced into thecombustion chamber 17. Since the hot exhaust gas is introduced into thecombustion chamber 17, the temperature inside the combustion chamber 17increases. Thus, it becomes advantageous to stabilize the SPCCIcombustion. Note that the intake-side electric S-VT 23 and theexhaust-side electric S-VT 24 may be provided with a negative overlapperiod where both the intake valve 21 and the exhaust valve 22 areclosed.

Moreover, the swirl generating part forms the strong swirl flow insidethe combustion chamber 17. The swirl ratio is four or higher, forexample. The swirl control valve 56 is fully closed or at a givenopening (closed to some extent). As described above, since the intakeport 18 is the tumble port, an inclined swirl flow having a tumblecomponent and a swirl component is formed in the combustion chamber 17.

The injector 6 injects fuel into the combustion chamber 17 a pluralityof times during the intake stroke (reference numerals 6011, 6012, 6021,6022, 6031, and 6032). The mixture gas is stratified by the multiplefuel injections and the swirl flow inside the combustion chamber 17.

The fuel concentration of the mixture gas in the central part of thecombustion chamber 17 is denser or richer than the fuel concentration inthe outer circumferential part. For example, the air-fuel ratio (A/F) ofthe mixture gas in the central part is 20 or higher and 30 or lower, andthe A/F of the mixture gas in the outer circumferential part is 35 orhigher. Note that the value of the A/F is a value when the mixture gasis ignited, and the same applies to the following description. Since theA/F of the mixture gas near the ignition plug 25 is set 20 or higher and30 or lower, generation of raw NO_(x) during the SI combustion can bereduced. Moreover, since the A/F of the mixture gas in the outercircumferential part is set to 35 or higher, the CI combustionstabilizes.

The A/F of the mixture gas is leaner than the stoichiometric air fuelratio throughout the combustion chamber 17 (i.e., excess air ratio λ>1).For example, the A/F of the mixture gas is 30 or higher throughout thecombustion chamber 17. Thus, the generation of raw NO_(x) can be reducedto improve the emission performance.

When the engine load is low (i.e., in the operating state 601), theinjector 6 performs the first injection 6011 in the first half of anintake stroke, and performs the second injection 6012 in the second halfof the intake stroke. The first half of the intake stroke may be a firsthalf of an intake stroke when the intake stroke is equally divided intotwo, and the second half of the intake stroke may be the rest. Moreover,an injection amount ratio of the first injection 6011 to the secondinjection 6012 may be 9:1, for example.

In the operating state 602 where the engine load is higher, the injector6 initiates the second injection 6022 which is performed in the secondhalf of an intake stroke at a timing advanced from the second injection6012 in the operating state 601. By advancing the second injection 6022,the mixture gas inside the combustion chamber 17 becomes morehomogeneous. The injection amount ratio of the first injection 6021 tothe second injection 6022 may be 7:3 to 8:2, for example.

In the operating state 603 where the engine load is even higher, theinjector 6 initiates the second injection 6032 which is performed in thesecond half of an intake stroke at a timing further advanced from thesecond injection 6022 in the operating state 602. By further advancingthe second injection 6032, the mixture gas inside the combustion chamber17 becomes further homogeneous. The injection amount ratio of the firstinjection 6031 to the second injection 6032 may be 6:4, for example.

After the fuel injection is finished, the ignition plug 25 ignites themixture gas in the central part of the combustion chamber 17 at a giventiming before a compression top dead center (reference numerals 6013,6023, and 6033). The ignition timing may be during a final stage of thecompression stroke. The compression stroke may be equally divided intothree, an initial stage, a middle stage, and a final stage, and thisfinale stage may be used as the final stage of the compression strokedescribed above.

As described above, since the mixture gas in the central part has therelatively high fuel concentration, the ignitability improves and the SIcombustion by flame propagation stabilizes. By the SI combustion beingstabilized, the CI combustion begins at a suitable timing. Thus, thecontrollability in the CI combustion improves in the SPCCI combustion.Further, the generation of combustion noise is reduced. Moreover, sincethe A/F of the mixture gas is made leaner than the stoichiometric airfuel ratio to perform the SPCCI combustion, fuel efficiency of theengine 1 can be significantly improved. Note that the low load area A1corresponds to Layer 3 described later. Layer 3 extends to the lightload operating range and includes a minimum load operating state.

(Middle-to-High Load Area)

When the engine 1 operates in the middle-to-high load area, the engine 1also performs the SPCCI combustion, similar to the low load area.

The reference numeral 604 in FIG. 8 indicates, in the middle-to-highload area, fuel injection timings (reference numerals 6041 and 6042), anignition timing (reference numeral 6043), and a combustion waveform(reference numeral 6044), when the engine 1 operates in the operatingstate 604 in the middle load area A2. The reference numeral 605indicates a fuel injection timing (reference numeral 6051), an ignitiontiming (reference numeral 6052), and a combustion waveform (referencenumeral 6053), when the engine 1 operates in the operating state 605 inthe high-load low-speed area A4.

The EGR system 55 introduces the EGR gas into the combustion chamber 17.For example, the intake-side electric S-VT 23 and the exhaust-sideelectric S-VT 24 are provided with a positive overlap period where boththe intake valve 21 and the exhaust valve 22 are opened near an exhausttop dead center. Internal EGR gas is introduced into the combustionchamber 17. Moreover, the EGR system 55 introduces the exhaust gascooled by the EGR cooler 53 into the combustion chamber 17 through theEGR passage 52. That is, the external EGR gas with a lower temperaturethan the internal EGR gas is introduced into the combustion chamber 17.The external EGR gas adjusts the temperature inside the combustionchamber 17 to a suitable temperature. The EGR system 55 reduces theamount of the EGR gas as the engine load increases. The EGR system 55may not recirculate the EGR gas containing the internal EGR gas and theexternal EGR gas during the full load.

Moreover, in the middle load area A2 and the high-load middle-speed areaA3, the swirl control valve 56 is fully closed or at a given opening(closed to some extent). In the combustion chamber 17, the strong swirlflow with the swirl ratio of four or higher is formed. On the otherhand, in the high-load low-speed area A4, the swirl control valve 56 isopen.

The air-fuel ratio (A/F) of the mixture gas is the stoichiometric airfuel ratio (A/F≈14.7:1) throughout the combustion chamber 17. Since thethree-way catalysts 511 and 513 purify the exhaust gas discharged fromthe combustion chamber 17, the emission performance of the engine 1 isimproved. The A/F of the mixture gas may be set within a purificationwindow of the three-way catalyst. The excess air ratio λ of the mixturegas may be 1.0±0.2. Note that when the engine 1 operates in thehigh-load middle-speed area A3 including the full load (i.e., themaximum load), the A/F of the mixture gas may be set at thestoichiometric air fuel ratio or richer than the stoichiometric air fuelratio (i.e., the excess air ratio λ of the mixture gas is λ≤1)throughout the combustion chamber 17.

Since the EGR gas is introduced into the combustion chamber 17, the G/Fwhich is a weight ratio of the entire gas to the fuel in the combustionchamber 17 becomes leaner than the stoichiometric air fuel ratio. TheG/F of the mixture gas may be 18:1 or higher. Thus, a generation of aso-called “knock” is avoided. The G/F may be set 18:1 or higher and 30:1or lower. Alternatively, the G/F may be set 18:1 or higher and 50:1 orlower.

When the engine 1 operates in the operating state 604, the injector 6performs a plurality of fuel injections (reference numerals 6041 and6042) during an intake stroke. The injector 6 may perform the firstinjection 6041 in the first half of the intake stroke and the secondinjection 6042 in the second half of the intake stroke.

Moreover, when the engine 1 operates in the operating state 605, theinjector 6 injects fuel in an intake stroke (reference numeral 6051).

The ignition plug 25 ignites the mixture gas at a given timing near acompression top dead center after the fuel is injected (referencenumerals 6043 and 6052). The ignition plug 25 may ignite the mixture gasbefore the compression top dead center when the engine 1 operates in theoperating state 604 (reference numeral 6043). The ignition plug 25 mayignite the mixture gas after the compression top dead center when theengine 1 operates in the operating state 605 (reference numeral 6052).

Since the A/F of the mixture gas is set to the stoichiometric air fuelratio and the SPCCI combustion is performed, the exhaust gas dischargedfrom the combustion chamber 17 can be purified using the three-waycatalysts 511 and 513. Moreover, the fuel efficiency of the engine 1improves by introducing the EGR gas into the combustion chamber 17 andmaking the mixture gas leaner. Note that the middle-to-high load areasA2, A3, and A4 correspond to Layer 2 described later. Layer 2 extends tothe high load area and includes the maximum load operating state.

(Operation of Supercharger)

Here, as illustrated in the map 504 of FIG. 7, the supercharger 44 isOFF in part of the low load area A1 and part of the middle load area A2(see S/C OFF). In detail, the supercharger 44 is OFF in an area on thelower engine speed side in the low load area A1. In an area on thehigher speed side in the low load area A1, the supercharger 44 is ON inorder to secure a required intake filling amount for the increased speedof the engine 1. Moreover, the supercharger 44 is OFF in a partial areaon the lower load and lower engine speed side in the middle load areaA2. In the area on the higher load side in the middle load area A2, thesupercharger 44 is ON in order to secure a required intake fillingamount for the increased fuel injection amount. Moreover, thesupercharger 44 is ON also in the area on the higher speed side in themiddle load area A2.

Note that in each area of the high-load middle-speed area A3, thehigh-load low-speed area A4, and the high speed area A5, thesupercharger 44 is entirely ON (see S/C ON).

(High Speed Area)

When the speed of the engine 1 increases, a time required for changingthe crank angle by 1° becomes shorter. Thus, it becomes difficult tostratify the mixture gas inside the combustion chamber 17. When thespeed of the engine 1 increases, it also becomes difficult to performthe SPCCI combustion.

Therefore, the engine 1 performs not the SPCCI combustion but the SIcombustion when the engine 1 operates in the high speed area A5. Notethat the high speed area A5 extends entirely in the load direction fromthe low load to the high load.

The reference numeral 606 of FIG. 8 indicates a fuel injection timing(reference numeral 6061), an ignition timing (reference numeral 6062),and a combustion waveform (reference numeral 6063), when the engine 1operates in the high speed area A5 in the operating state 606 where theload is high.

The EGR system 55 introduces the EGR gas into the combustion chamber 17.The EGR system 55 reduces the amount of the EGR gas as the loadincreases. The EGR system 55 may not recirculate the EGR gas during fullload.

The swirl control valve 56 is fully opened. No swirl flow occurs in thecombustion chamber 17, but only a tumble flow occurs. By fully openingthe swirl control valve 56, it is possible to increase the fillingefficiency, and reduce the pumping loss.

Fundamentally, the air-fuel ratio (A/F) of the mixture gas is thestoichiometric air fuel ratio (A/F≈14.7:1) throughout the combustionchamber 17. The excess air ratio λ of the mixture gas may be set to1.0±0.2. Note that the excess air ratio λ of the mixture gas may belower than 1 when the engine 1 operates near full load.

The injector 6 starts the fuel injection during an intake stroke. Theinjector 6 injects the fuel all at once (reference numeral 6061). Bystarting the fuel injection in the intake stroke, the homogeneous orsubstantially homogeneous mixture gas is formed inside the combustionchamber 17. Moreover, since a longer vaporizing time of the fuel can besecured, unburnt fuel loss can also be reduced.

After the fuel injection is finished, the ignition plug 25 ignites themixture gas at a suitable timing before a compression top dead center(reference numeral 6062).

(Layer Structure of Map)

As illustrated in FIG. 9, the maps 501, 502, and 503 of the engine 1illustrated in FIG. 6 are comprised of a combination of three layers,Layer 1, Layer 2, and Layer 3.

Layer 1 is a layer used as a base layer. Layer 1 extends throughout theoperating range of the engine 1. Layer 1 corresponds to the entire thirdmap 503.

Layer 2 is a layer which is superimposed on Layer 1. Layer 2 correspondsto part of the operating range of the engine 1. For example, Layer 2corresponds to the low-to-middle speed area B1, B2, and B3 of the secondmap 502.

Layer 3 is a layer which is superimposed on Layer 2. Layer 3 correspondsto the low load area A1 of the first map 501.

Layer 1, Layer 2, and Layer 3 are selected according to the walltemperature of the combustion chamber 17 and the intake air temperature.

When the wall temperature of the combustion chamber 17 is higher than agiven first wall temperature (e.g., 80° C.) and the intake airtemperature is higher than a given first intake air temperature (e.g.,50° C.), Layer 1, Layer 2, and Layer 3 are selected, and the first map501 is formed by piling up Layer 1, Layer 2, and Layer 3. In the lowload area A1 in the first map 501, the top Layer 3 therein becomeseffective, in the middle-to-high load areas A2, A3, and A4, the topLayer 2 therein becomes effective, and in the high speed area A5, Layer1 becomes effective.

When the wall temperature of the combustion chamber 17 is lower than thegiven first wall temperature and higher than a given second walltemperature (e.g., 30° C.), and the intake air temperature is lower thanthe given first intake air temperature and higher than a given secondintake air temperature (e.g., 25° C.), Layer 1 and Layer 2 are selected.By superimposing the Layer 1 and Layer 2, the second map 502 is formed.The low-to-middle speed area B1, B2, and B3 in second map 502, the topLayer 2 therein becomes effective, and in the high speed area B4, Layer1 becomes effective.

When the wall temperature of the combustion chamber 17 is lower than thegiven second wall temperature and the intake air temperature is lowerthan the given second intake air temperature, only Layer 1 is selectedto form the third map 503.

Note that the wall temperature of the combustion chamber 17 may bereplaced, for example, by temperature of the coolant of the engine 1measured by the water temperature sensor SW10. Alternatively, the walltemperature of the combustion chamber 17 may be estimated based on thetemperature of the coolant, or other measurements. The intake airtemperature is measurable, for example, by the third intake-airtemperature sensor SW17 which measures the temperature inside the surgetank 42. Alternatively, the temperature of the intake air introducedinto the combustion chamber 17 may be estimated based on various kindsof measurements.

As described above, the SPCCI combustion is performed by generating thestrong swirl flow inside the combustion chamber 17. Since the flamepropagates along the wall of the combustion chamber 17 during the SIcombustion, the flame propagation of the SI combustion is influenced bythe wall temperature. If the wall temperature is low, the flame of theSI combustion is cooled to delay the timing of compression-ignition.

Since the CI combustion of the SPCCI combustion is performed in the areafrom the outer circumferential part to the central part of thecombustion chamber 17, it is influenced by the temperature in thecentral part of the combustion chamber 17. If the temperature in thecentral part is low, the CI combustion becomes unstable. The temperaturein the central part of the combustion chamber 17 depends on thetemperature of the intake air introduced into the combustion chamber 17.That is, when the intake air temperature is higher, the temperature inthe central part of the combustion chamber 17 becomes higher, and whenthe intake air temperature is lower, the temperature in the central partbecomes lower.

When the wall temperature of the combustion chamber 17 is lower than thegiven second wall temperature and the intake air temperature is lowerthan the given second intake air temperature, a stable SPCCI combustioncannot be performed. Thus, only Layer 1 which performs the SI combustionis selected, and the ECU 10 operates the engine 1 based on the third map503. By the engine 1 performing the SI combustion in the entireoperating range, the combustion stability can be secured.

When the wall temperature of the combustion chamber 17 is higher thanthe given second wall temperature and the intake air temperature ishigher than the given second intake air temperature, the stable SPCCIcombustion of the mixture gas having substantially stoichiometric airfuel ratio (i.e., λ=1) can be carried out. Thus, in addition to Layer 1,Layer 2 is selected, and the ECU 10 operates the engine 1 based on thesecond map 502. By the engine 1 performing the SPCCI combustion in thepart of the operating ranges, the fuel efficiency of the engine 1improves.

When the wall temperature of the combustion chamber 17 is higher thanthe given first wall temperature and the intake air temperature ishigher than the given first the intake air temperature, the stable SPCCIcombustion of the mixture gas leaner than the stoichiometric air fuelratio can be carried out. Thus, in addition to Layer 1 and Layer 2,Layer 3 is selected, and the ECU 10 operates the engine 1 based on thefirst map 501. By the engine 1 performing the SPCCI combustion of thelean mixture gas in the part of the operating ranges, the fuelefficiency of the engine 1 further improves.

Next, one example of control related to the layer selection of the mapexecuted by the ECU 10 will be described with reference to a flowchartof FIG. 10. First, at Step S1 after the control is started, the ECU 10reads the signals of the sensors SW1-SW17. At the following Step S2, theECU 10 determines whether the wall temperature of the combustion chamber17 is 30° C. or higher and the intake air temperature is 25° C. orhigher. If the determination at Step S2 is YES, the control shifts theprocess to Step S3, and on the other hand, if NO, the control shifts theprocess to Step S5. The ECU 10 selects only Layer 1 at Step S5. The ECU10 operates the engine 1 based on the third map 503. The control thenreturns the process.

At Step S3, the ECU 10 determines whether the wall temperature of thecombustion chamber 17 is 80° C. or higher and the intake air temperatureis 50° C. or higher. If the determination at Step S3 is YES, the controlshifts the process to Step S4, and on the other hand, if NO, the controlshifts the process to Step S6.

The ECU 10 selects Layer 1 and Layer 2 at Step S6. The ECU 10 operatesthe engine 1 based on the second map 502. The control then returns theprocess.

The ECU 10 selects Layer 1, Layer 2, and Layer 3 at Step S4. The ECU 10operates the engine 1 based on the first map 501. The control thenreturns the process.

(Valve Timing of Intake Valve and Exhaust Valve)

FIG. 11 illustrates one example of a change in an open timing IVO of theintake valve 21 when the ECU 10 controls the intake-side electric S-VT23 according to the control logic set for Layer 2. An upper figure ofFIG. 11 (i.e., a graph 1101) illustrates a change of the open timing IVOof the intake valve 21 (vertical axis) versus the engine load(horizontal axis). The solid line corresponds to a case where the speedof the engine 1 is a relatively low first engine speed, and a brokenline corresponds to a case where the speed of the engine 1 is arelatively high second engine speed (first engine speed<second enginespeed).

A lower figure of FIG. 11 (i.e., a graph 1102) illustrates a change ofthe open timing IVO of the intake valve 21 (vertical axis) versus thespeed of the engine 1 (horizontal axis). The solid line corresponds to acase where the engine load is a relatively low first load, and thebroken line corresponds to a case where the engine load is a relativelyhigh second load (first load <second load).

In the graph 1101 and the graph 1102, the open timing IVO of the intakevalve 21 is advanced as it goes upward and the positive overlap periodwhere both the intake valve 21 and the exhaust valve 22 open becomeslonger. Therefore, the amount of the EGR gas introduced into thecombustion chamber 17 increases.

In Layer 2, the engine 1 operates with the A/F of the mixture gas at thestoichiometric air fuel ratio or the substantially stoichiometric airfuel ratio, and the G/F leaner than the stoichiometric air fuel ratio.When the engine load is low, the fuel supply amount decreases. Asillustrated in the graph 1101, when the engine load is low, the ECU 10sets the open timing IVO of the intake valve 21 at a timing on theretard side. Thus, the amount of the EGR gas introduced into thecombustion chamber 17 is regulated to secure the combustion stability.

Since the fuel supply amount increases when the engine load increases,the combustion stability improves. The ECU 10 sets the open timing ofthe intake valve 21 at a timing on the advance side. The pumping loss ofthe engine 1 can be lowered by increasing the amount of the EGR gasintroduced into the combustion chamber 17.

When the engine load further increases, the temperature inside thecombustion chamber 17 further increases. Then, the amount of theinternal EGR gas is reduced and the amount of the external EGR gas isincreased so that the temperature inside the combustion chamber 17 doesnot become too high. Therefore, the ECU 10 sets the open timing of theintake valve 21 again at a timing on the retard side.

When the engine load further increases and the supercharger 44 startsboosting, the ECU 10 sets the open timing of the intake valve 21 againat a timing on the advance side. Since the positive overlap period whereboth the intake valve 21 and the exhaust valve 22 open is provided, theresidual gas in the combustion chamber 17 can be purged.

Note that when the engine speed is high and low, the tendency of thechange in the open timing of the intake valve 21 is almost the same.

As illustrated in the graph 1102, when the engine speed is low, the flowinside the combustion chamber 17 becomes weaker. Since the combustionstability falls, the amount of the EGR gas introduced into thecombustion chamber 17 is regulated. The ECU 10 sets the open timing ofthe intake valve 21 at a timing on the retard side.

Since the flow inside the combustion chamber 17 becomes strong when theengine speed increases, the amount of the EGR gas introduced into thecombustion chamber 17 can be increased. The ECU 10 sets the open timingof the intake valve 21 at a timing on the advance side.

When the engine speed further increases, the ECU 10 sets the open timingof the intake valve 21 at a timing on the retard side according to theengine speed. Thus, the amount of gas introduced into the combustionchamber 17 is maximized.

FIG. 12 illustrates one example of a change in the open timing IVO ofthe intake valve 21, a close timing EVC of the exhaust valve 22, and anoverlap period O/L of the intake valve 21 and the exhaust valve 22, whenthe ECU 10 controls the intake-side electric S-VT 23 and theexhaust-side electric S-VT 24 according to the control logic set forLayer 3.

An upper figure of FIG. 12 (i.e., a graph 1201) illustrates a change inthe open timing IVO of the intake valve 21 (vertical axis) versus theengine load (horizontal axis). The solid line corresponds to a casewhere the engine speed is a relatively low third engine speed, and thebroken line corresponds to a case where the engine speed is relativelyhigh fourth engine speed (third engine speed<fourth engine speed).

A middle figure of FIG. 12 (i.e., graph 1202) illustrates a change inthe close timing EVC of the exhaust valve 22 (vertical axis) versus theengine load (horizontal axis). The solid line corresponds to a casewhere the engine speed is at the third engine speed, and the broken linecorresponds to a case where the engine speed is at the fourth enginespeed.

A lower figure of FIG. 12 (i.e., graph 1203) illustrates a change in theoverlap period O/L of the intake valve 21 and the exhaust valve 22(vertical axis) versus the engine load (horizontal axis). The solid linecorresponds to a case where the engine speed is at the third enginespeed, and the broken line corresponds to a case where the engine speedis at the fourth engine speed.

In Layer 3, the engine 1 operates by carrying out the SPCCI combustionof the mixture gas with the A/F leaner than the stoichiometric air fuelratio. When the engine load is low, the fuel supply amount decreases.Thus, the ECU 10 regulates the amount of gas introduced into thecombustion chamber 17 so that the A/F of the mixture gas does not becometoo low. As illustrated in the graph 1201, the ECU 10 sets the opentiming IVO of the intake valve 21 at a timing on the retard side of anexhaust top dead center. The close timing of the intake valve 21 becomesafter an intake bottom dead center, so-called “late close”. Moreover,when the engine load is low, the ECU 10 regulates the amount of theinternal EGR gas introduced into the combustion chamber 17. Asillustrated in the graph 1202, the ECU 10 sets the close timing EVC ofthe exhaust valve 22 at a timing on the advance side. The close timingEVC of the exhaust valve 22 approaches an exhaust top dead center.

Since the fuel supply amount increases when the engine load increases,the ECU 10 does not regulate the amount of gas introduced into thecombustion chamber 17. Moreover, in order to stabilize the SPCCIcombustion of the mixture gas leaner than the stoichiometric air fuelratio, the ECU 10 increases the amount of the internal EGR gasintroduced into the combustion chamber 17. The ECU 10 sets the opentiming IVO of the intake valve 21 at a timing on the advance side of anexhaust top dead center. Moreover, the ECU 10 sets the close timing EVCof the exhaust valve 22 at a timing on the retard side of the exhausttop dead center. As a result, as illustrated in the graph 1203, when theengine load increases, the overlap period where both the intake valve 21and the exhaust valve 22 open becomes longer.

When the engine load further increases, the ECU 10 reduces the amount ofthe internal EGR gas introduced into the combustion chamber 17 so thatthe temperature inside the combustion chamber 17 does not become toohigh. The ECU 10 brings the close timing EVC of the exhaust valve 22closer to an exhaust top dead center. Thus, the overlap period of theintake valve 21 and the exhaust valve 22 becomes shorter. Moreover, whenthe engine load is high and the engine speed is high, the ECU 10 setsthe open timing of the intake valve 21 on the retard side more than whenthe engine speed is low. Thus, the amount of gas introduced into thecombustion chamber 17 is maximized.

Note that in the low load area surrounded by a one-dot chain line inFIG. 12, the engine 1 may perform a reduced-cylinder operation in orderto improve fuel efficiency. When performing the reduced-cylinderoperation, the amount of gas and the internal EGR gas introduced intothe combustion chamber 17 are not regulated. As illustrated by a two-dotchain line in the graphs 1201 and 1202, the ECU 10 may set the opentiming of the intake valve 21 at a timing on the advance side and theclose timing of the exhaust valve 22 at a timing on the retard side.

(Engine Control Logic)

FIG. 13 is a flowchart illustrating the control logic of the engine 1.The ECU 10 operates the engine 1 according to the control logic storedin the memory 102. For example, the ECU 10 determines the operatingstate of the engine 1 based on the signals of the sensors SW1-SW17, andperforms calculations for adjusting properties in the combustion chamber17, the injection amount, the injection timing, and the ignition timingso that the combustion in the combustion chamber 17 becomes combustionat the SI ratio according to the operating state.

The ECU 10 first reads the signals of the sensors SW1-SW17 at Step S131.Subsequently, at Step S132, the ECU 10 determines the operating state ofthe engine 1 based on the signals of the sensors SW1-SW17, and sets atarget SI ratio (i.e., a target heat amount ratio). The target SI ratiois set according to the operating state of the engine 1.

FIG. 14 schematically illustrates one example of a setting of the targetSI ratio. When the engine load is low, the ECU 10 sets the target SIratio low, and on the other hand, when the engine load is high, it setsthe target SI ratio high. When the engine load is low, both thereduction of combustion noise and the improvement in fuel efficiency canbe achieved by increasing the ratio in the CI combustion to the SPCCIcombustion. When the engine load is high, it becomes advantageous forthe reduction of the combustion noise by increasing the ratio of the SIcombustion to the SPCCI combustion.

Returning to the flowchart of FIG. 13, the ECU 10 sets the targetin-cylinder properties for achieving the target SI ratio setting basedon the combustion model stored in the memory 102 at the following StepS133. For example, the ECU 10 sets a target temperature, a targetpressure, and target properties in the combustion chamber 17. At StepS134, the ECU 10 sets an opening of the EGR valve 54, an opening of thethrottle valve 43, an opening of the air bypass valve 48, an opening ofthe swirl control valve 56, and phase angles of the intake-side electricS-VT 23, and the exhaust-side electric S-VT 24 (i.e., a valve timing ofthe intake valve 21 and a valve timing of the exhaust valve 22), whichare required for achieving the target in-cylinder properties. The ECU 10sets the control amounts of these devices based on the map stored in thememory 102. The ECU 10 outputs signals to the EGR valve 54, the throttlevalve 43, the air bypass valve 48, the swirl control valve (SCV) 56, theintake-side electric S-VT 23, and the exhaust-side electric S-VT 24based on the control amount setting. By each device operating based onthe signal of the ECU 10, the properties in the combustion chamber 17become the target properties.

The ECU 10 further calculates predicted values or estimated values ofthe properties in the combustion chamber 17 based on the control amountsetting of each device. The predicted property value is a predictedvalue of the property in the combustion chamber 17 before the intakevalve 21 is closed. The predicted property value is used for setting ofthe fuel injection amount during the intake stroke as will be describedlater. The estimated property value is an estimated value of theproperty in the combustion chamber 17 after the intake valve 21 isclosed. The estimated property value is used for setting of the fuelinjection amount during a compression stroke, and setting of theignition timing, as will be described later. The estimated propertyvalue is also used for calculation of a property error of comparisonwith an actual combustion state.

At Step S135, the ECU 10 sets a fuel injection amount in the intakestroke based on the predicted property value. When performing a dividedinjection during the intake stroke, the ECU 10 sets the injection amountof each injection. Note that when fuel is not injected in the intakestroke, the injection amount of fuel is zero. At Step S136, the ECU 10outputs a signal to the injector 6 so that the injector 6 injects fuelinto the combustion chamber 17 at given injection timing(s).

At Step S137, the ECU 10 sets a fuel injection amount in a compressionstroke based on the estimated property value and the injection result ofthe fuel in the intake stroke. Note that when fuel is not injected inthe compression stroke, the injection amount of fuel is zero. At StepS138, the ECU 10 outputs a signal to the injector 6 so that the injector6 injects fuel into the combustion chamber 17 at an injection timingbased on the preset map.

At Step S139, the ECU 10 sets an ignition timing based on the estimatedproperty value and the injection result of the fuel in the compressionstroke. At Step S1310, the ECU 10 outputs a signal to the ignition plug25 so that the ignition plug 25 ignites the mixture gas inside thecombustion chamber 17 at the set ignition timing.

By the ignition plug 25 igniting the mixture gas, the SI combustion orthe SPCCI combustion is performed inside the combustion chamber 17. AtStep S1311, the ECU 10 reads the change in the pressure inside thecombustion chamber 17 measured by the pressure indicating sensor SW6,and based on the change, the ECU 10 determines a combustion state of themixture gas inside the combustion chamber 17. At Step S1312, the ECU 10compares the measurement result of the combustion state with theestimated property values estimated at Step S134, and calculates anerror between the estimated properties value and the actual properties.The calculated error is used for the estimation at Step S134 in thesubsequent cycles. The ECU 10 adjusts the openings of the throttle valve43, the EGR valve 54, the swirl control valve 56, and/or the air bypassvalve 48, and the phase angles of the intake-side electric S-VT 23 andthe exhaust-side electric S-VT 24 so as to eliminate the property error.Thus, the amount of the fresh air and the EGR gas introduced into thecombustion chamber 17 are adjusted.

If the ECU 10 estimates that the temperature inside the combustionchamber 17 will be lower than the target temperature again based on theestimated property values, it advances, at Step S138, the injectiontiming in the compression stroke more than the injection timing based onthe map so that the advancing of the ignition timing becomes possible.On the other hand, if ECU 10 estimates that the temperature inside thecombustion chamber 17 will be higher than the target temperature basedon the estimated property values, it retards, at Step S138, theinjection timing in the compression stroke more than the injectiontiming based on the map so that the retarding of the ignition timingbecomes possible.

That is, if the temperature inside the combustion chamber 17 is low, aself-ignition timing θ_(CI) (see FIG. 5) of the unburnt mixture gas isdelayed after the SI combustion begins by jump-spark ignition, whichdeviates the SI ratio from the target SI ratio. In this case, anincrease in the unburnt fuel and a fall in emission performance will becaused.

Thus, when the ECU 10 estimates that the temperature inside thecombustion chamber 17 will be lower than the target temperature, itadvances the injection timing, and advances the ignition timing at StepS1310. Since the heat release which is sufficient for the SI combustionbecomes possible by an earlier start of the SI combustion, it canprevent that the self-ignition timing θ_(CI) of the unburnt mixture gasis delayed when the temperature inside the combustion chamber 17 is low.As a result, the SI ratio approaches the target SI ratio.

On the other hand, when the temperature inside the combustion chamber 17is high, the unburnt mixture gas will carry out the self-ignitionshortly after the SI combustion begins by the jump-spark ignition, whichdeviates the SI ratio from the target SI ratio. In this case, combustionnoise increases.

Thus, when the ECU 10 estimates that the temperature inside thecombustion chamber 17 will be higher than the target temperature, itretards the injection timing, and retards the ignition timing at StepS1310. Since the start of the SI combustion is delayed, when thetemperature inside the combustion chamber 17 is high, the self-ignitiontiming θ_(CI) of the unburnt mixture gas becoming early can beprevented. As a result, the SI ratio approaches the target SI ratio.

The control logic of the engine 1 is configured to adjust the SI ratioby using a property setting device including the throttle valve 43, theEGR valve 54, the air bypass valve 48, the swirl control valve 56, theintake-side electric S-VT 23, and the exhaust-side electric S-VT 24. Byadjusting the properties in the combustion chamber 17, a roughadjustment of the SI ratio is possible. The control logic of the engine1 is also configured to adjust the SI ratio by adjusting the injectiontiming and the ignition timing of the fuel. By adjusting the injectiontiming and the ignition timing, a difference between the cylinders canbe corrected, and a fine adjustment of the self-ignition timing can beperformed, for example. By adjusting the SI ratio by two steps, theengine 1 can achieve the target SPCCI combustion accuratelycorresponding to the operating state.

The ECU 10 also outputs signals to at least the EGR system 55 and theignition plug 25 so that the actual SI ratio by combustion becomes thetarget SI ratio. Moreover, as described above, when the engine load ishigh, since the ECU 10 makes the target SI ratio higher than when theload is low, it then outputs the signals to at least the EGR system 55and the ignition plug 25 when the engine load is high so that the SIratio becomes higher than when the load is low.

Note that the control of the engine 1 executed by the ECU 10 is notlimited to the control logic based on the combustion model describedabove.

(Method of Controlling Engine)

When controlling the engine 1 for performing the SPCCI combustiondescribed above, the parameter related to the control amount of eachdevice is set. For example, if the device is the ignition plug 25,ignition energy and the ignition timing corresponding to the operatingstate of the engine 1 are set. If the device is the intake-side electricS-VT 23, the valve timing of the intake valve 21 corresponding to theoperating state of the engine 1 is set. Below, the setting of the valvetiming of the intake valve 21 in the method of controlling the engine 1will be described with reference to the drawings.

(1) Controlling Method According to Close Timing of Intake Valve

In the engine 1 for performing the SPCCI combustion, in order to reducecombustion noise and achieve the stable SPCCI combustion, the presentinventors found out that it was necessary to adjust the temperatureinside the combustion chamber 17 to a suitable temperature at a starttiming in the CI combustion (θ_(CI): see FIG. 5). That is, when thetemperature inside the combustion chamber 17 is low, the ignitability inthe CI combustion decreases. When the temperature inside the combustionchamber 17 is high, combustion noise increases.

The temperature inside the combustion chamber 17 at the start timingθ_(CI) in the CI combustion is mainly related to the effectivecompression ratio of the engine 1. The effective compression ratio ofthe engine 1 is determined by the geometric compression ratio ε and theclose timing IVC of the intake valve 21. In order to put the engine forperforming the SPCCI combustion in practical use, the present inventorsnewly found that a suitable IVC range existed within a range of thegeometric compression ratio ε where the SPCCI combustion may occur. Thetechnology disclosed herein is novel in that a given relation was foundout between the geometric compression ratio ε and the close timing IVCof the intake valve 21, where the given relation is required in order toput in practical use the engine which performs the unique combustionmode that is the SPCCI combustion. In addition, a relational expressionbetween the geometric compression ratio ε and the close timing IVC ofthe intake valve 21, which will be described later, is also novel.

Further, the technology disclosed herein is also novel in that, whencontrolling the engine 1 for performing the SPCCI combustion, the closetiming IVC of the intake valve 21 is set based on the relation betweenthe geometric compression ratio ε and the close timing IVC of the intakevalve 21.

The engine 1 switches between “Layer 2” where the SPCCI combustion isperformed with the A/F of the mixture gas being made at thestoichiometric air fuel ratio or richer than the stoichiometric air fuelratio and the G/F being made leaner than the stoichiometric air fuelratio, and “Layer 3” where the SPCCI combustion is performed with theA/F of the mixture gas being made leaner than the stoichiometric airfuel ratio. The relation between the geometric compression ratio ε andthe close timing IVC of the intake valve 21 in Layer 2 differs from therelation between the geometric compression ratio ε and the close timingIVC of the intake valve 21 in Layer 3.

(1-1) Relation Between Geometric Compression Ratio and Close Timing ofIntake Valve in Layer

FIG. 15 illustrates the characteristic of the SPCCI combustion. Forexample, FIG. 15 illustrates the occurring range of the SPCCI combustionagainst the EGR ratio (horizontal axis) in a case where the engine 1carries out the SPCCI combustion of the mixture gas of which the A/F isthe stoichiometric air fuel ratio and the G/F is leaner than thestoichiometric air fuel ratio, similar to Layer 2. The vertical axis inthis graph is a crank angle corresponding to a combustion center ofgravity, where the combustion center of gravity advances as it goesupward in this graph.

The occurring range of the SPCCI combustion is illustrated by a hatchedrange in this graph. The occurring range of the SPCCI combustion islocated between a line of “advancing limit” and a line of “retardinglimit.” If the combustion center of gravity advances beyond the“advancing limit” line, combustion becomes abnormal, which means thatthe SPCCI combustion does not occur. Similarly, if the combustion centerof gravity retards beyond the “retarding limit” line, self-ignition doesnot occur, which means that the SPCCI combustion does not occur.

The one-dot chain line in this graph illustrates the combustion centerof gravity of combustion corresponding to a MBT (Minimum advance forBest Torque). Here, the combustion center of gravity of combustioncorresponding to the MBT is simply referred to as “MBT.” The MBT isadvanced as the EGR ratio increases.

In terms of the improvement in fuel efficiency, it is desirable to bringthe combustion center of gravity of the SPCCI combustion closer to theMBT. As the EGR ratio increases, the occurring range of the SPCCIcombustion is advanced, but the interval between the “advancing limit”line and the “retarding limit” line becomes narrower, which narrows theoccurring range of the SPCCI combustion.

When the engine load is low (when “light load”), the occurring range ofthe SPCCI combustion is on the advance side. Therefore, as illustratedby a both-ends arrow in FIG. 15, the SPCCI combustion corresponding tothe MBT can be achieved by adjusting the EGR ratio within a certainwidth, and adjusting the combustion center of gravity to the advanceside or the retard side.

When the engine load increases, the amount of air introduced into thecombustion chamber 17 must be increased corresponding to the increase inthe fuel supply amount. When the EGR ratio is increased corresponding tothe increase in the air amount, a large amount of the EGR gas must beintroduced into the combustion chamber 17. However, because of the limitof the supercharging capability of the supercharger 44, it is difficultto introduce both a large amount of air and a large amount of the EGRgas into the combustion chamber 17. Therefore, when the engine loadincreases, the occurring range of the SPCCI combustion becomes on theretard side. If bringing the combustion center of gravity of the SPCCIcombustion closest to the MBT when the engine load is high, the SPCCIcombustion must be performed at a point Y in FIG. 15. The point Ycorresponds to the operating state of the engine 1 with the maximum loadwhere the SPCCI combustion of the mixture gas with the A/F being at thestoichiometric air fuel ratio is possible. When the engine 1 operateswith the maximum load, it is difficult to achieve the SPCCI combustioncorresponding to the MBT by adjusting the EGR ratio or the combustioncenter of gravity.

The engine 1 operates over a wide operating range from the low load tothe high load in Layer 2. In Layer 2, the state where the engine 1operates with the maximum load corresponds to the operating state at thelimit where the SPCCI combustion can occur. In Layer 2, upon determiningthe relation between ε and IVC so that the temperature of the combustionchamber 17 at the start timing (θ_(CI)) in the CI combustion becomes thegiven temperature, it is necessary to set the relation based on thetemperature inside the combustion chamber 17 when the engine 1 operateswith the maximum load.

In order to obtain the temperature inside the combustion chamber 17 whenthe engine 1 operates with the maximum load, the present inventorsperformed the SPCCI combustion in an actual engine 1, and usedmeasurements acquired from the engine 1. For example, the presentinventors measured various parameters when the engine 1 operates withthe maximum load in Layer 2, and estimated an actual temperature insidethe combustion chamber 17 at θ_(CI) based on the measured parameter. Thepresent inventors used an average value of a plurality of estimatedtemperatures as a reference temperature Tth1. If the temperature of thecombustion chamber 17 at θ_(CI) is the reference temperature Tth1, theSPCCI combustion can be achieved in Layer 2.

Here, as described above, θ_(CI) an be adjusted by adjusting theignition timing in the SPCCI combustion. However, while the engine 1operates with the maximum load in Layer 2, it becomes impossible toadjust θ_(CI) even if the ignition timing is adjusted when thetemperature inside the combustion chamber 17 at θ_(CI) exceeds thereference temperature Tth1. On the other hand, if the temperature insidethe combustion chamber 17 at θ_(CI) is the reference temperature Tth1 orlower, the ignition timing is adjusted (e.g., the ignition timing isadvanced) to raise the temperature inside the combustion chamber 17 atθ_(CI) to the reference temperature Tth1, thereby achieving the SPCCIcombustion.

Therefore, in order to achieve the SPCCI combustion in Layer 2, therelation between ε and IVC is required so that the temperature of thecombustion chamber 17 at θ_(CI) does not exceed the referencetemperature Tth1.

Thus, as conceptually illustrated in FIG. 16, the present inventorsperformed an estimation of the temperature of the combustion chamber 17at θ_(CI) by using a model of the engine 1, while changing the values ofthe geometric compression ratio ε and the close timing IVC of the intakevalve 21 in a matrix comprised of the two parameters of ε and IVC. Acombination of and IVC where the temperature of the combustion chamber17 becomes the reference temperature Tth1 or lower can achieve the SPCCIcombustion in Layer 2. As illustrated in FIG. 16, when the geometriccompression ratio ε is high and the close timing IVC of the intake valve21 approaches an intake bottom dead center, the temperature of thecombustion chamber 17 during the CI combustion exceeds the referencetemperature Tth1.

Note that FIG. 16 illustrates the matrix where IVC is set after theintake bottom dead center. Although illustration is omitted, the presentinventors also acquired a combination of and IVC where the temperaturebecomes the reference temperature Tth1 or lower, by performing anestimation of the temperature of the combustion chamber 17 at θ_(CI),similarly for the matrix where IVC is set before the intake bottom deadcenter.

A graph 1701 of the upper figure in FIG. 17 illustrates approximations(I) and (II) calculated based on the combination of ε and IVC. Thehorizontal axis of the graph 1701 is the geometric compression ratio ε,and the vertical axis is the close timing IVC (deg.aBDC) of the intakevalve 21. Although illustration is omitted, the present inventorsplotted on the graph 1701 the combination of ε and IVC where thetemperature becomes the reference temperature Tth1 or lower, anddetermined the approximations (I) and (II) based on the plots.

The graph 1701 corresponds to a case where the engine speed is 2,000rpm. The approximations (I) and (II) are as follows.IVC =−0.4288ε²+31.518ε−379.88   Approximation (I)IVC=−1.9163ε²−89.935ε+974.94   Approximation (II)

In the graph 1701, the combination of ε and IVC on the left side of theapproximations (I) and (II) and ε=17, the temperature of the combustionchamber 17 during the CI combustion becomes the reference temperatureTth1 or lower. In this combination, it is possible to carry out theSPCCI combustion of the mixture gas with the A/F being thestoichiometric air fuel ratio and the G/F being leaner than thestoichiometric air fuel ratio.

The relation between ε and IVC described above is a relation based onthe maximum temperature of the combustion chamber 17 in Layer 2.

On the other hand, in Layer 2, also while the engine 1 operates with thelight load, the relation between ε and IVC must be defined so that thetemperature of the combustion chamber 17 becomes the given temperature.

The temperature of the combustion chamber 17 when the SPCCI combustionis performed is a result of two pressure buildups of a pressure buildupby the compression work of the piston 3 in a compression stroke, and thepressure buildup caused by the heat generation of the SI combustion. Thecompression work of the piston 3 is determined by the effectivecompression ratio. If the effective compression ratio is too low, thepressure buildup by the compression work of the piston 3 decreases.Unless the pressure buildup, which is caused by the heat generation ofthe SI combustion after the flame propagation in the SPCCI combustionprogresses, increases considerably, the in-cylinder temperature cannotbe raised to an ignition temperature. As a result, since the amount ofthe mixture gas which is ignited by the compressed self-ignition islittle, and most of the mixture gas burn by the flame propagation, thecombustion period becomes longer and fuel efficiency falls. In order tostabilize the CI combustion in the SPCCI combustion and maximize fuelefficiency, it is necessary to maintain the effective compression ratioabove a certain value. Therefore, the relation between ε and IVC must bedetermined accordingly.

Similarly, the present inventors measured various parameters when theactual engine 1 operates with the light load, and estimated an actualtemperature inside the combustion chamber 17 at θ_(CI) based on themeasured parameters. The present inventors used an average value of aplurality of estimated temperatures as a reference temperature Tth2.

If the temperature inside the combustion chamber 17 at θ_(CI) is thereference temperature Tth2 or higher while the engine 1 operates withthe light load, the SPCCI combustion can be achieved by delaying theignition timing. However, since the temperature of the combustionchamber 17 is too low if the temperature at θ_(CI) is lower than thereference temperature Tth2, the SPCCI combustion cannot be achieved evenif the ignition timing is advanced.

Thus, in order to achieve the SPCCI combustion in Layer 2, a relationbetween ε and IVC is required so that the temperature of the combustionchamber 17 at θ_(CI) becomes the reference temperature Tth2 or higher.

The present inventors performed an estimation of the temperature of thecombustion chamber 17 during the CI combustion by using a model of theengine 1, while changing the values of the geometric compression ratio εand the close timing IVC of the intake valve 21 in a matrix comprised oftwo parameters of ε and IVC, similarly to the matrix illustrated in FIG.16. In this matrix, the combination of ε and IVC where the temperatureof the combustion chamber 17 becomes the reference temperature Tth2 orhigher can achieve the SPCCI combustion in Layer 2.

In the graph 1701 of FIG. 17, approximations (III) and (IV) calculatedbased on the combination of ε and IVC where the temperature becomes thereference temperature Tth2 or higher are also illustrated. Theapproximations (III) and (IV) are as follows.IVC=−0.4234ε²+22.926ε−167.84   Approximation (III)IVC=0.4234ε²−22.926ε+207.84   Approximation (IV)

In the graph 1701, the combination of ε and IVC on the right side of theapproximations (III) and (IV), the temperature of the combustion chamber17 during the CI combustion becomes the reference temperature Tth2 orhigher. In this combination, the SPCCI combustion of the mixture gaswith the A/F being the stoichiometric air fuel ratio and the G/F beingleaner than the stoichiometric air fuel ratio is possible.

As seen in FIG. 17, the relation between ε and IVC is substantiallyvertically symmetrical with respect to IVC=about 20 deg.aBDC. IVC=20deg.aBDC corresponds to a close timing (i.e., the best IVC) at which theamount of gas introduced into the combustion chamber 17 becomes themaximum when the engine speed is 2,000 rpm. Moreover, IVC=120 deg.aBDCis a retarding limit of the close timing IVC of the intake valve 21, andIVC=−80 deg.aBDC is an advancing limit of the close timing IVC of theintake valve 21.

The combination of ε and IVC within a range surrounded by theapproximations (I), (II), (III), and (IV) in FIG. 17 is a combinationwhich can put in practical use the engine 1 for performing the SPCCIcombustion in Layer 2. In other words, the combination of ε and IVCoutside this range cannot put in practical use the engine 1 forperforming the SPCCI combustion in Layer 2.

The ECU 10 has to determine IVC within the ε-IVC valid range hatched inFIG. 17, upon determining the close timing IVC of the intake valve 21when the engine 1 operates in Layer 2.

For example, if the geometric compression ratio ε is set as 10≤ε<17, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.4234ε²−22.926ε+207.84≤IVC≤−0.4234ε²+22.926ε−167.84   (1)

Moreover, if the geometric compression ratio ε is set as 17≤ε<20, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.4288ε²+31.518ε−379.88≤IVC≤−0.4234ε²+22.926ε−167.84   (2)or0.4234ε²−22.926ε+207.84≤IVC≤1.9163ε²−89.935ε+974.94   (3)

Further, if the geometric compression ratio ε is set as 20≤ε≤30, the ECU10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.4288ε²+31.518ε−379.88≤IVC≤120   (4)or−80≤IVC≤1.9163ε²−89.935ε+974.94   (5)

By setting the close timing IVC of the intake valve 21 based on therelational expressions (1) to (5), the SPCCI combustion of the mixturegas with the A/F being the stoichiometric air fuel ratio or richer thanthe stoichiometric air fuel ratio and the G/F being leaner than thestoichiometric air fuel ratio is achieved. Note that the close timingIVC is set for each operating state which is determined by the load andthe engine speed in Layer 2. The example illustrated by the solid linein FIG. 17 is the ε-IVC valid range when the engine speed is 2,000 rpm,as described above. If the engine speed changes, the ε-IVC valid rangealso changes. As the engine speed increases, the best IVC is retarded.

For example, when the engine speed is 3,000 rpm, the best IVC is about22 deg.aBDC. As illustrated by broken lines in FIG. 17, the ε-IVC validrange when the engine speed is 3,000 rpm can be obtained by parallellytranslating the ε-IVC valid range when the engine speed is 2,000 rpm tothe retard side by about 2 degrees.

Therefore, if the geometric compression ratio ε is set as 10≤ε<17, whenthe engine speed is 3,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.0.4234ε²−22.926ε+209.84≤IVC≤−0.4234ε²+22.926ε−165.84   (1⁻¹)

Moreover, if the geometric compression ratio ε is set as 17≤ε<20, whenthe engine speed is 3,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.−0.4288ε²+31.518ε−377.88≤IVC≤−0.4234ε²+22.926ε−165.84   (2⁻¹)or0.4234ε²−22.926ε+209.84≤IVC≤1.9163ε²−89.935ε+976.94   (3⁻¹)

Further, if the geometric compression ratio ε is set as 20≤ε≤<30, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.4288ε²+31.518ε−377.88≤IVC≤120   (4⁻¹)or−80≤IVC≤1.9163ε²−87.935ε+976.94   (5⁻¹)

Moreover, when the engine speed is 4,000 rpm, the best IVC is about 28deg.aBDC. As illustrated by one-dot chain lines in FIG. 17, the ε-IVCvalid range when the engine speed is 4,000 rpm is obtained by parallellytranslating the ε-IVC valid range when the engine speed is 2,000 rpm tothe retard side by about 8 degrees.

Therefore, if the geometric compression ratio ε is set as 10≤ε<17, whenthe engine speed is 4,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.0.4234ε²−22.926ε+215.84≤IVC≤−0.4234ε²+22.926ε−159.84   (1⁻²)

Moreover, if the geometric compression ratio ε is set as 17≤ε<20, whenthe engine speed is 4,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.−0.4288ε²+31.518ε−371.88≤IVC≤−0.4234ε²+22.926ε−159.84   (2⁻²)or0.4234ε²−22.926ε+215.84≤IVC≤1.9163ε²−89.935ε+982.94   (3⁻²)

Further, if the geometric compression ratio ε is set as 20≤ε≤30, the ECU10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.4288ε²+31.518ε−371.88≤IVC≤120   (4⁻²)or−80≤IVC≤1.9163ε²−89.935ε+982.94   (5⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is set as the following,C=3.3×10⁻ NE ³−1.0×10⁻⁶ NE ³+7.0×10⁻⁴ NEthe relational expression of ε and IVC in Layer 2 can be expressed asfollows. If the geometric compression ratio ε is 10≤ε<17,0.4234ε²−22.926ε+207.84+C≤IVC≤−0.4234ε²+22.926ε−167.84+C   (1⁻³)If the geometric compression ratio ε is 17≤ε<20,0.4288ε²+31.518ε−379.88+C≤IVC≤−0.4234ε²+22.926ε−167.84+C   (2⁻³)or0.4234ε²−22.926ε+207.84+C≤IVC≤1.9163ε²−89.935ε+974.94+C   (3⁻³)If the geometric compression ratio ε is 20≤ε≤30,−0.4288ε²+31.518ε−379.88+C≤IVC≤120   (4⁻³)or−80≤IVC≤1.9163ε²−89.935ε+974.94+C   (5⁻³)The ECU 10 determines the close timing IVC based on the ε-IVC validrange determined for every engine speed of the engine 1. As a result,the ECU 10 can set the valve timing of the intake valve 21 in Layer 2 asillustrated in FIG. 11.(1-2) Change in ε-IVC Valid Range by Difference of Octane Number

The graph 1701 in FIG. 17 is a relation between ε and IVC when the fuelis high octane fuel (octane number is about 96). A graph 1702illustrated in the lower figure is a relation between ε and IVC when thefuel is low octane fuel (octane number is about 91). According to theexamination of the present inventors, when the fuel was the low octanefuel, it was found that the ε-IVC valid range shifts by 1.3 compressionratios toward the lower compression ratio from the ε-IVC valid range ofthe high octane fuel.

Accordingly, upon determining the close timing IVC in the engine 1 ofthe low octane fuel, if the geometric compression ratio ε is set as10≤ε<15.7, when the engine speed is 2,000 rpm, the ECU 10 determines theclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.4234ε²−21.826ε+178.75≤IVC≤−0.4234ε²+21.826ε−138.75   (6)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.5603ε²+34.859ε−377.22≤IVC≤−0.4234ε²+21.826ε−138.75   (7)or0.4234ε²−21.826ε+178.75≤IVC≤1.9211ε²−85.076ε+862.01   (8)

Further, if the geometric compression ratio ε is set as 18.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 2,000 rpm, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.5603ε²+34.859ε−377.22≤IVC≤120   (9)or−80≤IVC≤1.9211ε²−85.076ε+862.01   (10)

The cross-hatched range in the graph 1702 of FIG. 17 is an overlappingrange of the ε-IVC valid range of the high octane fuel and the ε-IVCvalid range of the low octane fuel. The ECU 10 can set the control logicwhich suits both the engine 1 using the high octane fuel and the engine1 using the low octane fuel, if IVC is determined within the overlappingrange of the two valid ranges. Even if the octane number of the fueldiffers for every destination of this product, the ECU 10 cancollectively use the engine control logic.

Note that although illustration is omitted, the ε-IVC valid range isparallelly translated to the retard side also in the engine 1 of the lowoctane fuel, when the engine speed increases. If the geometriccompression ratio ε is set as 10≤ε<15.7 in the engine 1 of the lowoctane fuel, when the engine speed is 3,000 rpm, the ECU 10 determinesthe close timing IVC (deg.aBDC) so that the following expression issatisfied.0.4234ε²−21.826ε+180.75≤IVC≤−0.4234ε²+21.826ε−136.75   (6⁻¹)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.5603ε²+34.859ε−375.22≤IVC≤−0.4234ε²+21.826ε−136.75   (7⁻¹)or0.4234ε²−21.826ε+180.75≤IVC≤1.9211ε²−85.076ε+864.01   (8⁻¹)

Further, if the geometric compression ratio ε is set as 18.7≤ε<30 in theengine 1 of the low octane fuel, when the engine speed is 3,000 rpm, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.5603ε²+34.859ε−375.22≤IVC≤120   (9⁻¹)or−80≤IVC≤1.9211ε²−85.076ε+864.01   (10⁻¹)

Moreover, if the geometric compression ratio ε is set as 10≤ε<15.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.0.4234ε²−21.826ε+186.75≤IVC≤−0.4234ε²+21.826ε−130.75   (6⁻²)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.5603ε²+34.859ε−369.22≤IVC≤−0.4234ε²+21.826ε−130.75   (7⁻²)or0.4234ε²−21.826ε+186.75≤IVC≤1.9211ε²−85.076ε+870.01   (8⁻²)

Further, if the geometric compression ratio ε is set as 18.7≤ε<30 in theengine 1 of the low octane fuel, when the engine speed is 4,000 rpm, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.5603ε²+34.859ε−369.22≤IVC≤120   (9⁻²)or−80≤IVC≤1.9211ε²−77.076ε+870.01   (10⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 2 in the engine 1 of the low octane fuel can beexpressed as follows.

If the geometric compression ratio ε is 10≤ε≤15.7,0.4234ε²−21.826ε+178.75+C≤IVC≤−0.4234ε²+21.826ε−138.75+C   (6⁻³)If the geometric compression ratio ε is 15.7≤ε<18.7,−0.5603ε²+34.859ε−377.22+C≤IVC≤−0.4234ε²+21.826ε−138.75+C   (7⁻³)or0.4234ε²−21.826ε+178.75+C≤IVC≤1.9211ε²−85.076ε+862.01+C   (8⁻³)If the geometric compression ratio ε is 18.7≤ε≤30,−0.5603ε²+34.859ε−377.22+C≤IVC≤120   (9⁻³)or−80≤IVC≤1.9211ε²−85.076ε+862.01+C   (10⁻³)(1-3) Relation Between Geometric Compression Ratio and Close Timing ofIntake Valve in Layer 3

FIG. 18 illustrates the characteristic of the SPCCI combustion when theengine 1 carries out the SPCCI combustion of the mixture gas where theA/F is leaner than the stoichiometric air fuel ratio, similar to Layer3. FIG. 18 illustrates a range where the SPCCI combustion is stable withrespect to the G/F (horizontal axis). The vertical axis in this graph isa crank angle corresponding to a combustion center of gravity, and thecombustion center of gravity is advanced as it goes upward in thisgraph.

The range where the SPCCI combustion is stabilized is a range surroundedby a curve in this graph. When the engine load is low, the range wherethe SPCCI combustion is stabilized is located at upper left in FIG. 18.When the engine load increases, the range where the SPCCI combustion isstabilized moves downwardly in FIG. 18.

FIG. 18 also illustrates a range where discharge of raw NO_(x) can bereduced. The range where the discharge of raw NO_(x) can be reduced islocated at lower right in FIG. 18. This range has a triangular shape inFIG. 18. If the A/F is leaner than the stoichiometric air fuel ratio,raw NO_(x) cannot be purified by the three-way catalyst. In Layer 3, theengine 1 must satisfy both securing of the stability of the SPCCIcombustion and reduction of the discharge of raw NO_(x).

As seen in this graph, if the engine load is high, an overlapping areaof the range where the combustion stability is secured and the rangewhere the discharge of raw NO_(x) can be reduced increases. On the otherhand, if the engine load is low, the overlapping area of the range wherethe combustion stability is secured and the range where the discharge ofraw NO_(x) can be reduced decreases.

Regarding Layer 3, the state where the engine 1 operates with the lightload corresponds to a limit of the operating state where the SPCCIcombustion can occur. Regarding Layer 3, upon determining the relationbetween ε and IVC so that the temperature of the combustion chamber 17at the start timing (θ_(CI)) of the CI combustion becomes the giventemperature, it is necessary to be set based on the temperature insidethe combustion chamber 17 when the engine 1 operates with the lightload.

Similar to the above, using the actual engine 1, the present inventorsmeasured various parameters when operating the engine 1 with the lightload in Layer 3, and estimated an actual temperature inside thecombustion chamber 17 at θ_(CI) based on the measured parameters. Then,an average value of a plurality of estimated temperatures was determinedas a reference temperature Tth3. If the temperature of the combustionchamber 17 at θ_(CI) is the reference temperature Tth3, the SPCCIcombustion can be achieved in Layer 3. This reference temperature Tth3corresponds to the minimum temperature. If the temperature inside thecombustion chamber 17 at θ_(CI) is the reference temperature Tth3 orhigher while the engine 1 operates with the light load, the SPCCIcombustion can be achieved by delaying the ignition timing. However, ifthe temperature inside the combustion chamber 17 at θ_(CI) is lower thanthe reference temperature Tth3, SPCCI combustion cannot be achieved evenif the ignition timing is advanced.

Therefore, in order to achieve the stable SPCCI combustion in Layer 3,the relation between ε and IVC is required so that the temperatureinside the combustion chamber 17 at θ_(CI) becomes the referencetemperature Tth3 or higher.

Thus, as conceptually illustrated in FIG. 19, the present inventorsestimated the temperature of the combustion chamber 17 during the CIcombustion by using the model of the engine 1, while changing the valuesof IVC and O/L in a matrix of the close timing IVC of the intake valve21 and the overlap period O/L of the intake valve 21 and the exhaustvalve 22, for every geometric compression ratio ε (ε1, ε2 . . . ) (areference numeral 1901). As hatched in the matrix of the referencenumeral 1901, a combination of IVC and O/L where the temperature becomesthe reference temperature Tth3 or higher can achieve the SPCCIcombustion in Layer 3.

Moreover, in order to reduce the discharge of raw NO_(x), the G/F of themixture gas must be made a given value or higher. As illustrated by areference numeral 1902 in FIG. 19, the present inventors estimated theG/F by using the model of the engine 1, while changing the values of theclose timing IVC and the overlap period O/L in the matrix comprised ofthe two parameters of IVC and O/L. As oblique lines are drawn in thematrix of the reference numeral 1902, the combination of IVC and O/Lwhere the G/F becomes the given value or higher can reduce the dischargeof raw NO_(x).

Then, the present inventors determined the relation between ε and IVCwhich can achieve both the stability of the SPCCI combustion and thereduction of raw NO_(x) discharge, by overlapping the combination of IVCand O/L where the temperature becomes the reference temperature Tth3 orhigher which is illustrated by the reference numeral 1901, and thecombination of IVC and O/L where the G/F becomes the given value orhigher which is illustrated by the reference numeral 1902. That is, inthe matrix of a reference numeral 1903, a cross-hatched range is thecombination of ε and IVC which can achieve both the stability of theSPCCI combustion and the reduction of NO_(x) discharge.

Note that although illustration is omitted, the present inventorsacquired a combination of IVC and O/L where the temperature becomes thereference temperature Tth3 or higher and a combination of IVC and O/Lwhere the G/F becomes the given value or higher by estimating thetemperature of the combustion chamber 17 and the G/F during the CIcombustion, similarly for the matrix where the close timing of theintake valve 21 is set before the intake bottom dead center.

FIG. 20 illustrates approximations (V) and (VI) calculated based on thecombinations of ε and IVC. The horizontal axis in FIG. 20 is thegeometric compression ratio ε, and the vertical axis is the close timingIVC (deg. aBDC) of the intake valve 21.

An upper figure 2001 of FIG. 20 corresponds to a case when the enginespeed is 2,000 rpm. The approximations (V) and (VI) are as follows.IVC=−0.9949ε²+41.736ε−361.16   Approximation (V)IVC=0.9949ε²−41.736ε+401.16   Approximation (VI)

In FIG. 20, the combinations of ε and IVC on the right side of theapproximations (V) and (VI) have the temperature of the combustionchamber 17 during CI combustion becoming the reference temperature Tth3or higher, thereby achieving the SPCCI combustion of the mixture gaswith the A/F being leaner than the stoichiometric air fuel ratio.

In Layer 3, the relation between ε and IVC described above is a relationbased on the minimum temperature of the combustion chamber 17 which canachieve the SPCCI combustion when the engine 1 operates with the lightload.

On the other hand, if the temperature inside the combustion chamber 17is too high, the CI combustion begins before the start of the SIcombustion and the SPCCI combustion cannot be performed, regardless ofLayer 2 or Layer 3.

Here, the concept of the SPCCI combustion is such that, as describedabove, when the ignition plug 25 ignites the mixture gas, the mixturegas around the ignition plug 25 starts the SI combustion, and, afterthat, the surrounding mixture gas carries out the CI combustion. Fromexaminations by experiments, etc. which the present inventors conducteduntil now, it was found that the self-ignition of the mixture gasoccurred when the surrounding temperature of the mixture gas whichself-ignites exceeds a given reference temperature Tth4, and thisreference temperature Tth4 was not necessarily in agreement with a meantemperature of the entire combustion chamber 17. From this knowledge, ifthe mean temperature inside the combustion chamber 17 at a compressiontop dead center reaches the reference temperature Tth4, it is thoughtthat the CI combustion will begin before the SI combustion begins, andin this case, the SPCCI combustion cannot be performed.

Thus, the present inventors estimated the temperature of the combustionchamber 17 at the compression top dead center by using the model of theengine 1, while changing the values of the close timing IVC and theoverlap period O/L in the matrix of the close timing IVC of the intakevalve 21, and the overlap period O/L of the intake valve 21 and theexhaust valve 22, for every geometric compression ratio ε (ε1, ε2 . . .), similar to the matrix of the reference numeral 1901 of FIG. 19. Thecombination of IVC and O/L where the temperature inside the combustionchamber 17 at the compression top dead center exceeds the referencetemperature Tth4 cannot achieve the SPCCI combustion, but thecombination of IVC and O/L being the reference temperature Tth4 or lowercan achieve the SPCCI combustion.

FIG. 20 illustrates approximations (VII) and (VIII) calculated based onthe combination of ε and IVC where the temperature inside the combustionchamber 17 at the compression top dead center becomes the referencetemperature Tth4 or lower. The approximations (VII) and (VIII) are asfollows.IVC=−4.7481ε²+266.75ε−3671.2   Approximation (VII)andIVC=4.7481ε²−266.75ε+3711.2   Approximation (VIII)

In FIG. 20, the combination of ε and IVC on the left side theapproximations (VII) and (VIII) can avoid that the CI combustion beginsbefore the SI combustion, and achieves the SPCCI combustion.

As seen in FIG. 20, also in Layer 3, the relation between ε and IVC issubstantially vertically symmetrical with respect to IVC=20 deg.aBDC.Moreover, IVC=75 deg.aBDC is a retarding limit of the close timing ofthe intake valve 21 set in consideration of the amount of gas introducedinto the combustion chamber 17 when the engine 1 operates in Layer 3.Similarly, IVC=−35 deg.aBDC is an advancing limit of the close timing ofthe intake valve 21 set in consideration of the amount of gas introducedinto the combustion chamber 17.

When determining the close timing IVC of the intake valve 21 in the casewhere the engine 1 operates in Layer 3, the ECU 10 must determine IVCwithin the ε-IVC valid range surrounded by the approximations (V), (VI),(VII), and (VIII) in FIG. 20 (the hatched range in FIG. 20).

For example, if the geometric compression ratio ε is set as 10≤ε<20,when the engine speed is 2,000 rpm, the ECU 10 determines the closetiming IVC (deg.aBDC) so that the following expression is satisfied.0.9949ε²−41.736ε+401.16≤IVC≤−0.9949ε²+41.736ε−361.16   (11)

Moreover, if the geometric compression ratio ε is set as 20≤ε<25, whenthe engine speed is 2,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.−35≤IVC≤75   (12)

Further, if the geometric compression ratio ε is set as 25≤ε≤30, whenthe engine speed is 2,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.−4.7481ε²+266.75ε−3671.2≤IVC≤75   (13)or−35≤IVC≤4.7481ε²−266.75ε+3711.2   (14)

By setting the close timing IVC of the intake valve 21 based on therelational expressions (11) to (14), the SPCCI combustion is achieved ofthe mixture gas with a A/F leaner than the stoichiometric air fuelratio. Note that the close timing IVC is set for each operating statewhich is determined by the load and the engine speed in Layer 3.

The example illustrated by a solid line in FIG. 20 is the ε-IVC validrange when the engine speed is 2,000 rpm, as described above. If theengine speed changes, the ε-IVC valid range also changes. When theengine speed increases, the ε-IVC valid range is also parallellytranslated to the retard side, similarly in FIG. 20. Therefore, if thegeometric compression ratio is set as 10≤ε<20, when the engine speed is3,000 rpm (see a broken line), the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.0.9949ε²−41.736ε+403.16≤IVC≤−0.9949ε²+41.736ε−359.16   (11⁻¹)

If the geometric compression ratio ε is set as 20≤ε<25, when the enginespeed is 3,000 rpm, the ECU 10 determines the close timing IVC(deg.aBDC) so that the following expression is satisfied.−33≤IVC≤77   (12⁻¹)

Further, if the geometric compression ratio ε is set as 25≤ε≤30, whenthe engine speed is 3,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.−4.7481ε²+266.75ε−3669.2≤IVC≤77   (13⁻¹)or−33≤IVC≤4.7481ε²−266.75ε+3713.2   (14⁻¹)

If the geometric compression ratio ε is set as 10≤ε≤20, when the enginespeed is 4,000 rpm (see one-dot chain line), the ECU 10 determines theclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−41.736ε+409.16≤IVC≤−0.9949ε²+41.736ε−353.16   (11⁻²)

If the geometric compression ratio ε is set as 20≤ε≤25, when the enginespeed is 4,000 rpm, the ECU 10 determines the close timing IVC(deg.aBDC) so that the following expression is satisfied.−27≤IVC≤83   (12⁻²)

Further, if the geometric compression ratio ε is set as 25≤ε≤30, whenthe engine speed is 4,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.−4.7481ε²+266.75ε−3663.2≤IVC≤83   (13⁻²)or−27≤IVC≤4.7481ε²−266.75ε+3719.2   (14⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 3 can be expressed as follows. If the geometriccompression ratio ε is 10≤ε<20,0.9949ε²−41.736ε+401.16+C≤IVC≤−0.9949ε²+41.736ε−361.16+C   (11⁻³)If the geometric compression ratio ε is 20≤ε≤25,−35+C≤IVC≤75+C   (12⁻³)If the geometric compression ratio ε is 25≤ε≤30,−4.7481ε²+266.75ε−3671.2+C≤IVC≤75+C   (13⁻³)or−35+C≤IVC≤4.7481ε²−266.75ε+3711.2+C   (14⁻³)

The ECU 10 determines the close timing IVC based on the ε-IVC validrange determined for every engine speed of the engine 1. As a result,the ECU 10 can set the valve timing of the intake valve 21 in Layer 3 asillustrated in FIG. 12.

Moreover, a lower figure 2002 of FIG. 20 is a relation between ε and IVCwhen the fuel is the low octane fuel.

Upon determining the close timing IVC in the engine 1 of the low octanefuel, if the geometric compression ratio ε is set as 10≤ε<18.7, when theengine speed is 2,000 rpm, the ECU 10 determines the close timing IVC(deg.aBDC) so that the following expression is satisfied.0.9949ε²−39.149ε+348.59≤IVC≤−0.9949ε²+39.149ε−308.59   (15)

Moreover, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−35≤IVC≤75   (16)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 2,000 rpm, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−3.1298ε²+172.48ε−2300≤IVC≤75   (17)or−35≤IVC≤3.1298ε²−172.48ε+2340   (18)

The cross-hatched range in the lower figure 2002 of FIG. 20 is anoverlapping range of the ε-IVC valid range of the high octane fuel andthe ε-IVC valid range of the low octane fuel. Similar to the above, thecontrol logic which suits both the engine 1 using the high octane fueland the engine 1 using the low octane fuel is set by determining IVCwithin the overlapping range of the two occurring ranges.

Note that although illustration is omitted, if the engine speedincreases, the ε-IVC valid range is parallelly translated to the retardside also in the engine 1 of the low octane fuel. If the geometriccompression ratio ε is set as 10≤ε<18.7 in the engine 1 of the lowoctane fuel, when the engine speed is 3,000 rpm, the ECU 10 determinesthe close timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−39.149ε+350.59≤IVC≤−0.9949ε²+39.149ε−306.59   (15⁻¹)

Moreover, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−33≤IVC≤77   (16⁻¹)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 3,000 rpm, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−3.1298ε²+172.48ε−2298≤IVC≤77   (17⁻¹)or−33≤IVC≤3.1298ε²−172.48ε²+2342   (18⁻¹)

If the geometric compression ratio ε is set as 10≤ε<18.7 in the engine 1of the low octane fuel, when the engine speed is 4,000 rpm, the ECU 10determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.9949ε²−39.149ε+356.59≤IVC≤−0.9949ε²+39.149ε−300.59   (15⁻²)

Moreover, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−27≤IVC≤83   (16⁻²)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 4,000 rpm, theECU 10 determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−3.1298ε²+172.48ε−2292≤IVC≤83   (17⁻²)or−27≤IVC≤3.1298ε²−172.48ε+2348   (18⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 3, in the engine 1 of the low octane fuel can beexpressed as follows. If the geometric compression ratio ε is 10≤ε<18.7,0.9949ε²−39.149ε+348.59+C≤IVC≤−0.9949ε²+39.149ε−308.59+C   (15⁻³)If the geometric compression ratio ε is 18.7≤ε<23.7,−35+C≤IVC≤75+C   (16⁻³)If the geometric compression ratio ε is 23.7≤εε30,−3.1298ε²+172.48ε−2300+C≤IVC≤75+C   (17⁻³)or−35+C≤IVC≤3.1298ε²−172.48ε+2340+C   (18⁻³)(1-4) Relation Between Geometric Compression Ratio and Close Timing ofIntake Valve in Layers 2 and 3

FIG. 21 illustrates a relation between the geometric compression ratio εand the close timing IVC of the intake valve 21 where the SPCCIcombustion is possible in both Layer 2 and Layer 3. This relationalexpression is obtained from the ε-IVC valid range of FIG. 17 and theε-IVC valid range of FIG. 20.

When the ECU 10 selects Layer 3 according to the temperature, etc. ofthe engine 1, the low-load operating range of the engine 1 is switchedfrom Layer 2 to Layer 3. If the close timing IVC of the intake valve 21is set so that the SPCCI combustion is possible in both Layer 2 andLayer 3, it is possible to continuously perform the SPCCI combustioneven when the map of the engine 1 is switched from Layer 2 to Layer 3.

An upper figure 2101 of FIG. 21 is a relation between ε and IVC when thefuel is the high octane fuel. A lower figure 2102 is a relation betweenε and IVC when the fuel is the low octane fuel.

If the geometric compression ratio ε is set as 10≤ε<17 in the engine 1of the high octane fuel, when the engine speed is 2,000 rpm, the ECU 10determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.9949ε²−41.736ε+401.16≤IVC≤−0.9949ε²+41.736ε−361.16   (19)

If the geometric compression ratio ε is set as 17≤ε≤30 in the engine 1of the high octane fuel, when the engine speed is 2,000 rpm, the ECU 10determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.4288ε²+31.518ε−379.88≤IVC≤−0.9949ε²+41.736ε−361.16   (20)or0.9949ε²−41.736ε+401.16≤IVC≤1.9163ε²−89.935ε+974.94   (21)

By setting the close timing IVC of the intake valve 21 based on therelational expressions (19) to (21), the SPCCI combustion of the mixturegas with the A/F being leaner than the stoichiometric air fuel ratio canbe carried out, and the SPCCI combustion of the mixture gas with the A/Fbeing the stoichiometric air fuel ratio or richer than thestoichiometric air fuel ratio, and the G/F being leaner than thestoichiometric air fuel ratio can be carried out.

Note that the close timing IVC is set for each operating state which isdetermined by the load and the engine speed in Layer 2 and Layer 3.

As illustrated by a broken line, if the geometric compression ratio ε isset as 10≤ε<17 in the engine 1 of the high octane fuel, when the enginespeed is 3,000 rpm, the ECU 10 determines the close timing IVC(deg.aBDC) so that the following expression is satisfied.0.9949ε²−41.736ε+403.16≤IVC≤−0.9949ε²+41.736ε−359.16   (19⁻¹)

If the geometric compression ratio ε is set as 17≤ε≤30 in the engine 1of the high octane fuel, when the engine speed is 3,000 rpm, the ECU 10determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.4288ε²+31.518ε−377.88≤IVC≤−0.9949ε²+41.736ε−359.16   (20⁻¹)or0.9949ε²−41.736ε+403.16≤IVC≤1.9163ε²−89.935ε+976.94   (21⁻¹)

Similarly, as illustrated by a one-dot chain line, if the geometriccompression ratio ε is set as 10≤ε<17 in the engine 1 of the high octanefuel, when the engine speed is 4,000 rpm, the ECU 10 determines theclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−41.736ε+409.16≤IVC≤−0.9949ε²+41.736ε−353.16   (19⁻²)

If the geometric compression ratio ε is set as 17≤ε≤30 in the engine 1of the high octane fuel, when the engine speed is 4,000 rpm, the ECU 10determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.−0.4288ε²+31.518ε−371.88≤IVC≤−0.9949ε²+41.736ε−353.16   (20⁻²)or0.9949ε²−41.736ε+409.16≤IVC≤1.9163ε²−89.935ε+982.94   (21)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 2 and Layer 3 can be expressed as follows. If thegeometric compression ratio ε is 10≤ε<17,0.9949ε²−41.736ε+401.16+C≤IVC≤−0.9949ε²+41.736ε−361.16+C   (19⁻³)If the geometric compression ratio ε is 17≤ε≤30,−0.4288ε²+31.518ε−379.88+C≤IVC≤−0.9949ε²+41.736ε−361.16+C   (20⁻³)or0.9949ε²−41.736ε+401.16+C≤IVC≤1.9163ε²−89.935ε+974.94+C   (21⁻³)

Here, if the geometric compression ratio ε determined to be lower than17, the ECU 10 can determine IVC based on the relational expression(19⁻³). Since the selection range of IVC is wide, a degree of freedom inthe design becomes high.

Moreover, as illustrated in the lower figure 2102 of FIG. 21, if thegeometric compression ratio ε is set as 10≤ε<15.7 in the engine 1 of thelow octane fuel, when the engine speed is 2,000 rpm, the ECU 10determines the close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.9949ε²−39.149ε+348.59≤IVC≤−0.9949ε²+39.149ε−308.59   (22)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε≤30 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.5603ε²+34.859ε−377.22≤IVC≤−0.9949ε²+39.149ε−308.59   (23)or0.9949ε²−39.149ε+348.59≤IVC≤1.9211ε²−85.076ε+862.01   (24)

Although illustration is omitted, if the geometric compression ratio εis set as 10≤ε<15.7 in the engine 1 of the low octane fuel, when theengine speed is 3,000 rpm, the ECU 10 determines the close timing IVC(deg.aBDC) so that the following expression is satisfied.0.9949ε²−39.149ε+350.59≤IVC≤−0.9949ε²+39.149ε−306.59   (22⁻¹)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε≤30 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.5603ε²+34.859ε−375.22≤IVC≤−0.9949ε²+39.149ε−306.59   (23⁻¹)or0.9949ε²−39.149ε+350.59≤IVC≤1.9211ε²−85.076ε+864.01   (24⁻¹)

Although illustration is similarly omitted, if the geometric compressionratio ε is set as 10≤ε<15.7 in the engine 1 of the low octane fuel, whenthe engine speed is 4,000 rpm, the ECU 10 determines the close timingIVC (deg.aBDC) so that the following expression is satisfied.0.9949ε²−39.149ε+356.59≤IVC≤−0.9949ε²+39.149ε−300.59   (22⁻²)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε≤30 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the ECU 10 determines the close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.5603ε²+34.859ε−369.22≤IVC≤−0.9949ε²+39.149ε−300.59   (23⁻²)or0.9949ε²−39.149ε+356.59≤IVC≤1.9211ε²−85.076ε+870.01   (24⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 2 and Layer 3 in the engine 1 of the low octane fuelcan be expressed as follows. If the geometric compression ratio ε is10≤ε<15.7,0.9949ε²−39.149ε+348.59+C≤IVC≤−0.9949ε²+39.149ε−308.59+C   (22⁻³)If the geometric compression ratio ε is 15.7≤ε≤30,−0.5603ε²+34.859ε−377.22+C≤IVC≤−0.9949ε²+39.149ε−308.59+C   (23⁻³)or0.9949ε²−39.149ε+348.59+C≤IVC<1.9211ε²−85.076ε+862.01+C   (24⁻³)

Note that although illustration is omitted, the ECU 10 may determine IVCwithin an overlapping range of the ε-IVC valid range of the upper FIG.2101 and the ε-IVC valid range of the lower figure 2102 of FIG. 21.Similar to the above, the ECU 10 can set the control logic which suitsboth the engine 1 using the high octane fuel and the engine 1 using thelow octane fuel by determining IVC within the overlapping range of thetwo occurring ranges.

(1-5) Procedure of Method of Implementing Control Logic

Next, a procedure of the method of controlling the engine 1 forperforming the SPCCI combustion will be described with reference to aflowchart illustrated in FIG. 22. The ECU 10 can execute each step andstores information on the ε-IVC valid range illustrated in FIGS. 17, 20,and 21.

At Step S221 after the procedure starts, the ECU 10 first sets thegeometric compression ratio ε. The set value of the geometriccompression ratio ε may be inputted from an exterior device into the ECU10.

At the following Step S222, the ECU 10 sets the valve opening angle ofthe intake valve 21, and the valve opening angle of the exhaust valve22. This corresponds to determining the cam shapes of the intake valve21 and the exhaust valve 22. The set values of the valve opening anglesof the intake valve 21 and the exhaust valve 22 may be inputted from anexterior device into the ECU 10. Thus, a hardware configuration of theengine 1 can be identified at Steps S221 and S222.

At Step S223, the ECU 10 sets the operating state comprised of the loadand the engine speed, and at the following Step S224, the ECU 10 selectsIVC based on the stored ε-IVC valid range (FIGS. 17, 20, and 21).

Then, at Step S225, the computer determines whether the SPCCI combustioncan be achieved based on IVC set at Step S224. If the determination atStep S225 is YES, this procedure shifts to Step S226, and the ECU 10controls the engine 1 according to the control logic corresponding tothe engine operating state so that the SPCCI combustion is performed inthe operating state set at Step S223. On the other hand, if thedetermination at Step S225 is NO, this procedure shifts to Step S227,and the ECU 10 controls the engine 1 according to the control logiccorresponding to the engine operating state so that the SI combustion isperformed in the operating state set at Step S223. Note that at StepS227, the ECU 10 may again set the close timing IVC of the intake valve21 in consideration of performing the SI combustion.

As described above, the control device of the compression-ignitionengine disclosed herein can set the close timing IVC of the intake valve21 within the range where the relation between the engine geometriccompression ratio ε and the close timing IVC of the intake valve 21 issatisfied.

(2) Method of Controlling Supercharging Pressure

As described above, the close timing IVC of the intake valve 21 can beset based on the relation between the geometric compression ratio ε andthe close timing IVC of the intake valve 21.

As a result of further repeated and diligent examinations, the presentinventors further found that, in order to put in practical use theengine 1 for performing the SPCCI combustion with a lean G/F in Layer 2,another relation was required between a supercharging pressure P of theengine 1 and the close timing IVC of the intake valve 21. The presentinventors further found a control method based on such another relation,after the implementing method was utilized.

(2-1) Relation Between Supercharging Pressure and Close Timing of IntakeValve in Layer 2

As described above, the engine 1 operates, while making the A/F of themixture gas the stoichiometric air fuel ratio or a substantiallystoichiometric air fuel ratio in Layer 2, and making the G/F leaner thanthe stoichiometric air fuel ratio. Here, the G/F can be controlledthrough the amount of gas supplied into the combustion chamber 17, and acontrol factor for controlling the amount of gas includes thesupercharging pressure P of the supercharger 44, in addition to anoverlap period O/V.

For example, since the fuel supply amount is large when operating theengine 1 in the high-load operating state including the maximum loadoperating state, a larger amount of gas is required to be supplied intothe combustion chamber 17. Therefore, in order to maintain the G/F leanin Layer 2, it may be necessary to set the supercharging pressure P(especially, a target supercharging pressure of the supercharger 44) ata given pressure or higher. That is, in Layer 2, a lower limit of thesupercharging pressure P can be determined. Moreover, an actual amountof gas sent into the combustion chamber 17 from the intake passage 40among the gas supercharged by the supercharger 44 depends also on theclose timing IVC of the intake valve 21.

Therefore, as a condition for maintaining the G/F at lean in Layer 2, arelation between the lower limit of the supercharging pressure P and theclose timing IVC can be determined.

Therefore, the present inventors estimated the G/F by using the model ofthe engine 1 also in consideration of an inertia effect of air, whilechanging the values of the two parameters of the supercharging pressureP and the close timing IVC. Then, the present inventors acquired acombination of the lower limit of the supercharging pressure P and theclose timing IVC so that the G/F becomes the given value or higher.

FIG. 23 illustrates an approximation (A) calculated from the combinationof the lower limit of P and IVC. The horizontal axis of FIG. 23 is theclose timing IVC (deg. aBDC) of the intake valve 21, and the verticalaxis is the supercharging pressure P (kPa) of the engine 1. In FIG. 23,“E^(−n)” (n is an integer) indicates “10^(−n).” For example, “8E-11IVC⁶”means “8×10⁻¹¹×IVC⁶.”

An upper figure of FIG. 23 corresponds to a case where the engine speedof the engine 1 is 2000 rpm or higher. The approximation (A) correspondsto a solid line in FIG. 23.P=8.0×10⁻¹¹ IVC ⁶−1.0×10⁻⁸ IVC ⁵+3.0×10⁻⁷IVC⁴−4.0×10⁻⁶IVC³+0.0068IVC²−0.3209IVC+116.63   Approximation (A)

In FIG. 23, the combination of P and IVC above the approximation (A) canachieve the SPCCI combustion, while maintaining the G/F at lean, becausethe G/F becomes the given value or higher.

As seen in FIG. 23, the relation between P and IVC is substantiallylaterally symmetrical with respect to IVC=20 deg.aBDC. IVC=20 deg.aBDCcorresponds to a close timing (best IVC) at which the amount of gasintroduced into the combustion chamber 17 becomes the maximum.

That is, at IVC=20 deg.aBDC, a large amount of gas can be supplied intothe combustion chamber 17, even if the supercharging pressure P is setrelatively low. Therefore, the supercharging pressure P is the minimumat IVC=20 deg.aBDC. Moreover, in order to maintain the G/F at a givenvalue or higher, it is required to set the supercharging pressure Prelatively high as the G/F deviates from this minimum. Therefore, as IVCdeviates from 20 deg.aBDC, it is necessary to increase the superchargingpressure P.

Note that the close timing IVC is set for each operating state which isdefined by the load and the engine speed in Layer 2. The example of FIG.23 is a P-IVC valid range when the engine speed of the engine 1 is 2000rpm, as described above. If the engine speed of the engine 1 changes,the P-IVC valid range also changes. The ECU 10 determines the closetiming IVC based on the P-IVC valid range defined for every engine speedof the engine 1.

The supercharging pressure P described above may also be a targetsupercharging pressure of the supercharger 44 in Layer 2.

Moreover, it is generally thought that, as for the superchargingpressure P which can be set to the supercharger 44, an upper limit (aso-called supercharger limit) due to its hardware configuration exists.Therefore, in Layer 2, the upper limit can be provided to thesupercharging pressure P.

FIG. 23 also illustrates the approximation (B) indicative of the upperlimit of P. The approximation (B) corresponds to the reference characterP1 in FIG. 23, and in the illustrated example,P=150.   Approximation (B)

In FIG. 23, the combination of P and IVC below the approximation (B)permits the supercharging pressure P to be set within a range withoutexceeding the supercharger limit. Moreover, although the superchargerlimit indicated by the approximation (B) varies according to thehardware configuration of the supercharger 44, the limit may be set soas to at least exceed the minimum of the supercharging pressure P in theapproximation (A).

Moreover, other than the upper limit of P based on the approximation(B), an upper limit (second upper limit) of P in consideration of thefuel efficiency may be provided. Thus, the loss of work caused byexcessive supercharging pressure can be reduced, and this becomesadvantageous to improve the fuel efficiency of the engine. Althoughillustration is omitted, this second upper limit can be set to, forexample,P=250.

As illustrated in the above equation, the second upper limit can be sethigher than the upper limit illustrated in the approximation (B).

Moreover, in order to achieve the SPCCI combustion in Layer 2, therelation between ε and IVC is required in the first place so that thetemperature of the combustion chamber 17 at θCI becomes the referencetemperature Tth2 or higher. Such a requirement is satisfied by, forexample, the combination of ε and IVC on the right side of theapproximations (III) and (IV) in FIG. 17.

That is, in order to achieve the SPCCI combustion in Layer 2, IVC needsto fall within a given range. That is, an upper limit and a lower limitcan be set to IVC.

Thus, when estimating the G/F, the present inventors set the upper limitand the lower limit of IVC according to the value of the geometriccompression ratio ε.

In FIG. 23, an approximation (C) indicative of the upper limit of IVCand an approximation (D) indicative of the lower limit of IVC are alsoillustrated. The approximation (C) is determined based on theapproximation (III) in FIG. 17, and corresponds to a reference characterIr in FIG. 23. On the other hand, the approximation (D) is determinedbased on the approximation (IV) in FIG. 17, and corresponds to areference character Il in FIG. 23. In the illustrated example, theapproximations (C) and (D) are as follows.IVC=90   Approximation (C)IVC=−50   Approximation (D)

In FIG. 23, the combination of P and IVC on the left side of theapproximation (C) and on the right side of the approximation (D) canstably achieve SPCCI combustion.

Note that the approximations (C) and (D) in the upper figure of FIG. 23correspond to a case where the geometric compression ratio ε is set toaround 16 as can be seen in FIG. 17. If ε is set smaller than this, asillustrated in the lower figure of FIG. 23, the upper limit and thelower limit of IVC both approach IVC=20 deg.aBDC. Moreover, theapproximations (C) and (D) vary according to the octane number of thefuel.

Note that the approximation (A) in the lower figure of FIG. 23 is madethe same as that of the upper figure for the convenience ofillustration, but the details of the approximation (A) may vary in factaccording to the geometric compression ratio ε.

The combination of P and IVC within a range surrounded by theapproximations (A), (B), (C), and (D) in FIG. 23 is the combinationwhich can achieve the SPCCI combustion, while maintaining the G/F inLayer 2 at the given value or higher. In other words, the combination ofP and IVC outside this range cannot achieve the G/F-lean SPCCIcombustion in Layer 2.

The ECU 10 must determine IVC within the P-IVC valid range which ishatched in FIG. 23, when operating the engine 1 in Layer 2 anddetermining the close timing IVC of the intake valve 21.

Specifically, in the configuration example illustrated in the upperfigure of FIG. 23, the ECU 10 determines the close timing IVC (deg.aBDC)so that the supercharging pressure P (kPa) satisfies the followingexpression.P≥8.0×10⁻¹¹ IVC ⁶−1.0×10⁻⁸ IVC ⁵+3.0×10⁻⁷ IVC ⁴−4.0×10⁻⁶ IVC³+0.0068IVC²−0.3209IVC+116.63   (25)

Moreover, the ECU 10 determines the close timing IVC (deg.aBDC) so thatthe supercharging pressure P (kPa) satisfies the following expression.P≤150   (26)

Moreover, the ECU 10 determines the close timing IVC (deg.aBDC) so as tosatisfy the following expression.−50≤IVC≤90   (27)(2-2) Procedure of Control Method

Next, a procedure of the method of controlling the engine 1 forperforming the SPCCI combustion will be described with reference to aflowchart illustrated in FIG. 24. The ECU 10 can execute each of thefollowing steps. The ECU 10 stores information on the P-IVC valid rangeillustrated in FIG. 23.

At Step S241 after the procedure starts, the ECU 10 first sets thegeometric compression ratio ε. A preset value of the geometriccompression ratio ε may be inputted into the ECU 10 from an exteriordevice.

At the subsequent Step S242, the ECU 10 sets a valve opening angle ofthe intake valve 21, and a valve opening angle of the exhaust valve 22.This corresponds to determining cam shapes of the intake valve 21 andthe exhaust valve 22. The preset values of the valve opening angles ofthe intake valve 21 and the exhaust valve 22 may be inputted into theECU 10 from an exterior device. A hardware configuration of the engine 1can be specified at Steps S241 and S242.

At Step S243, the ECU 10 sets the operating state comprised of the loadand the engine speed of the engine 1, and at the subsequent Step S244,the ECU 10 sets a target supercharging pressure as the superchargingpressure P based on the operating state set at Step S243.

At the subsequent Step S245, the ECU 10 determines IVC based on theprestored P-IVC valid range (FIG. 23).

Then, at Step S246, the ECU 10 determines whether SPCCI combustion canbe achieved based on IVC set at Step S245. If the determination of StepS246 is YES, the procedure shifts to Step S247, where the ECU 10controls the engine 1 according to the control logic corresponding tothis operating state so that the SPCCI combustion is performed in theoperating state set at Step S243. On the other hand, if thedetermination of Step S246 is NO, the procedure shifts to Step S248,where the ECU 10 controls the engine 1 according to the control logiccorresponding to this operating state so that the SI combustion isperformed in the operating state set at Step S243. Note that at StepS245, the ECU 10 may again set the close timing IVC of the intake valve21 in consideration of performing the SI combustion.

As described above, in the control device of the compression-ignitionengine disclosed herein, the relation between the supercharging pressureP of the engine 1 and the close timing IVC of the intake valve 21 isdefined beforehand, and the ECU 10 can set the close timing IVC of theintake valve 21 within a range where this relation is satisfied.

(Other Embodiments)

Note that the application of the technology disclosed herein is notlimited to the engine 1 having the configuration described above. Theengine 1 may adopt various configurations.

For example, the engine 1 may be provided with a turbocharger, insteadof the mechanical supercharger 44.

It should be understood that the embodiments herein are illustrative andnot restrictive, since the scope of the invention is defined by theappended claims rather than by the description preceding them, and allchanges that fall within metes and bounds of the claims, or equivalenceof such metes and bounds thereof, are therefore intended to be embracedby the claims.

DESCRIPTION OF REFERENCE CHARACTERS

1 Engine

10 ECU (Control Part)

17 Combustion Chamber

23 Intake-side Electric S-VT (Variable Valve Operating Mechanism)

25 Ignition Plug (Ignition Part)

44 Supercharger

55 EGR System

6 Injector (Fuel Injection Part)

SW1 Airflow Sensor (Measurement Part)

SW2 First Intake-air Temperature Sensor (Measurement Part)

SW3 First Pressure Sensor (Measurement Part)

SW4 Second Intake-air Temperature Sensor (Measurement Part)

SW5 Second Pressure Sensor (Measurement Part)

SW6 Pressure Indicating Sensor (Measurement Part)

SW7 Exhaust Temperature Sensor (Measurement Part)

SW8 Linear O₂ Sensor (Measurement Part)

SW9 Lambda O₂ Sensor (Measurement Part)

SW10 Water Temperature Sensor (Measurement Part)

SW11 Crank Angle Sensor (Measurement Part)

SW12 Accelerator Opening Sensor (Measurement Part)

SW13 Intake Cam Angle Sensor (Measurement Part)

SW14 Exhaust Cam Angle Sensor (Measurement Part)

SW15 EGR Pressure Difference Sensor (Measurement Part)

SW16 Fuel Pressure Sensor (Measurement Part)

SW17 Third Intake-air Temperature Sensor (Measurement Part)

What is claimed is:
 1. A method of implementing control logic of acompression-ignition engine, the engine comprising: a fuel injectionpart configured to inject fuel to be supplied in a combustion chamber; avariable valve operating mechanism configured to change a valve timingof an intake valve; an ignition part configured to ignite a mixture gasinside the combustion chamber; a supercharger configured to boost gasintroduced into the combustion chamber; a measurement part configured tomeasure a parameter related to an operating state of the engine; and acontrol part configured to perform a calculation according to a controllogic corresponding to the operating state of the engine, in response tothe measurement of the measurement part, and output a signal to the fuelinjection part, the variable valve operating mechanism, the ignitionpart, and the supercharger, the control part outputting the signal tothe fuel injection part, the variable valve operating mechanism and thesupercharger so that a gas-fuel ratio (G/F) that is a weight ratio ofthe entire gas of the mixture gas inside the combustion chamber to thefuel becomes leaner than a stoichiometric air fuel ratio, and anair-fuel ratio (A/F) that is a weight ratio of air contained in themixture gas to the fuel becomes the stoichiometric air fuel ratio orricher than the stoichiometric air fuel ratio, while causing thesupercharger to boost, and outputting the signal to the ignition part sothat an unburnt mixture gas combusts by self-ignition after the ignitionpart ignites the mixture gas inside the combustion chamber, the methodof implementing the control logic, comprising the steps of: determininga supercharging pressure P by the supercharger; and determining thecontrol logic defining a close timing IVC of the intake valve, wherein,when determining the control logic, the close timing IVC (deg.aBDC) isdetermined so that the supercharging pressure P (kPa) satisfies thefollowing expression:P≥8.0×10⁻¹¹ IVC ⁶−1.0×10⁻⁸ IVC ⁵+3.0×10⁻⁷ IVC ⁴−4.0×10⁻⁶ IVC ³+0.0068IVC²−0.3209IVC+116.63.
 2. The method of claim 1, wherein the superchargingpressure (kPa) is determined so as to satisfy P≤150.
 3. The method ofclaim 1, wherein the control part determines a target superchargingpressure corresponding to the operating state of the engine, and whereinthe control part determines the close timing IVC (deg.aBDC) so that,when the target supercharging pressure is used as the superchargingpressure P (kPa), the supercharging pressure P (kPa) satisfies theexpression.
 4. The method of claim 1, wherein the close timing IVC ofthe intake valve changes as the operating state of the engine changes,and wherein the close timing IVC (deg.aBDC) is determined for eachoperating state so that the expression is satisfied.
 5. The method ofclaim 1, wherein the engine operates in a high-load operating state at agiven load or higher.
 6. The method of claim 5, wherein the engineoperates in a maximum load operating state.
 7. The method of claim 1,wherein a geometric compression ratio ε of the engine is set so as tosatisfy 10≤ε<21.
 8. The method of claim 1, wherein the engine isprovided with an exhaust gas recirculation (EGR) system configured tointroduce exhaust gas into the combustion chamber, and wherein thecontrol part outputs the signal to the EGR system and the ignition partso that a heat amount ratio used as an index related to a ratio of anamount of heat generated when the mixture gas inside the combustionchamber combusts by flame propagation to the entire amount of heatgenerated when the mixture gas combusts, becomes a target heat amountratio defined corresponding to the operating state of the engine.
 9. Themethod of claim 8, wherein the control part outputs a signal to the EGRsystem and the ignition part so that the heat amount ratio becomeshigher when the load of the engine is higher.
 10. A control device for acompression-ignition engine, the engine comprising: a fuel injectionpart configured to inject fuel to be supplied in a combustion chamber; avariable valve operating mechanism configured to change a valve timingof an intake valve; an ignition part configured to ignite a mixture gasinside the combustion chamber; a supercharger configured to boost gasintroduced into the combustion chamber; and a measurement partconfigured to measure a parameter related to an operating state of theengine, the control device comprising: a control part configured toperform a calculation according to a control logic corresponding to theoperating state of the engine, in response to the measurement of themeasurement part, and output a signal to the fuel injection part, thevariable valve operating mechanism, the ignition part, and thesupercharger, the control part outputting the signal to the fuelinjection part, the variable valve operating mechanism and thesupercharger so that a gas-fuel ratio (G/F) that is a weight ratio ofthe entire gas of the mixture gas inside the combustion chamber to thefuel becomes leaner than a stoichiometric air fuel ratio, and anair-fuel ratio (A/F) that is a weight ratio of air contained in themixture gas to the fuel becomes the stoichiometric air fuel ratio orricher than the stoichiometric air fuel ratio, while causing thesupercharger to boost, and outputting the signal to the ignition part sothat an unburnt mixture gas combusts by self-ignition after the ignitionpart ignites the mixture gas inside the combustion chamber, wherein thecontrol part determines a supercharging pressure P by the supercharger,and determines a close timing IVC of the intake valve, and wherein thecontrol part determines, according to the control logic, the closetiming IVC (deg.aBDC) so that the supercharging pressure P (kPa)satisfies the following expression:P≥8.0×10⁻¹¹ IVC ⁶−1.0×10⁻⁸ IVC ⁵+3.0×10⁻⁷ IVC ⁴−4.0×10⁻⁶ IVC ³+0.0068IVC²−0.3209IVC+116.63.
 11. The control device of claim 10, wherein thesupercharging pressure (kPa) is determined so as to satisfy P≤150. 12.The control device of claim 10, wherein the control part determines atarget supercharging pressure corresponding to the operating state ofthe engine, and wherein the control part determines the close timing IVC(deg.aBDC) so that, when the target supercharging pressure is used asthe supercharging pressure P (kPa), the supercharging pressure P (kPa)satisfies the expression.
 13. The control device of claim 10, whereinthe close timing IVC of the intake valve changes as the operating stateof the engine changes, and wherein the close timing IVC (deg.aBDC) isdetermined for each operating state so that the expression is satisfied.14. The control device of claim 10, wherein the engine operates in ahigh-load operating state at a given load or higher.
 15. The controldevice of claim 14, wherein the engine operates in a maximum loadoperating state.
 16. The control device of claim 10, wherein thegeometric compression ratio ε of the engine is set so as to satisfy10≤ε<21.
 17. The control device of claim 10, wherein the engine isprovided with an EGR system configured to introduce exhaust gas into thecombustion chamber, and wherein the control part outputs the signal tothe EGR system and the ignition part so that a heat amount ratio used asan index related to a ratio of an amount of heat generated when themixture gas inside the combustion chamber combusts by flame propagationto the entire amount of heat generated when the mixture gas combusts,becomes a target heat amount ratio defined corresponding to theoperating state of the engine.
 18. The control device of claim 17,wherein the control part outputs a signal to the EGR system and theignition part so that the heat amount ratio becomes higher when the loadof the engine is higher.